Automotive Transmissions: Fundamentals, Selection, Design and Application

Gisbert Lechner · Harald Naunheimer

Automotive Transmissions Fundamentals, Selection, Design and Application In Collaboration with Joachim Ryborz

With 370 Figures

Springer

erms and Symbols

Uv 'C

fP fPI f/J2 fPth (i)

XXI

Speed ratio, kinematic viscosity Density, angle of friction of the claws Direct stress Bending stress Fatigue strength Hertzian stresses Reference stress Torsional stress, torque increase with combustion engine Gear step, bending angle Basic step with progressive stepping Progression factor with progressive stepping Gear step with geometrical stepping . Angular velocity

Subscripts

o 1

2 1,2,3, ... A B C CG CS D E Ex F G H IS L L, Ll, L2 M MS N OS P PV

Q R Roll S Sch St T TC U V

Nominal or initial state Pinion (= small gear), input Wheel (= large gear), output At point 1,2,3, ... Offer, related to area, drive shaft, power train, moving off Demand, brake Clutch Constant gear Countershaft Duration, fatigue-resistant, deficit, opening, direct drive End Excess Vehicle, root, free-wheel Gearbox Adhesion, main, main gearbox, main shaft wheel, ring gear, high (= fast) Input shaft Air, load, low (= slow) At bearing point, at bearing point 1, 2 Engine, motor, model Main shaft Rear-mounted range-change unit Output shaft Pump, pump wheel, planetary step Pump test Transverse Reverse gear, roll, slip, friction, wheel, range-change unit, reactor Roll Status, system, splitter unit Pulsating (strength) Gradient Turbine wheel, drive Torque converter Circumference rtv~lt ~.1~OUIJ.~C~ splittci" u(~it, "/arlZttc: .lo~;s, t:-i8.! 4

,

J_.......- - - r - - - - - - - - - - - - - - ' - - - -

Professor Dr.-Ing. Gisbert Lechner

Dr.-Ing. Harald Naunheimer

Director IUMllule oi MaClline Components Drive Technology. CAD. Sealing Technology. Reliability Technology University of Stuttgart Pfaffenwaldring 9. 70569 Stuttgart Germany

Head of Project Engineering and Kesearch Vehicle Transmission Division Renk AG Gogginger StraBe 73. 86159 Augsburg Germany

Collaborator:

Translation Agency:

Dipl.-Ing. Joachim Ryborz Institute ofMachine Components Drive Technology. CAD. Sealing Technology. ReliabilityTechnology UniversityofStuttgart Pfaffenwaldring9. 70569 Stuttgart Germany

Ingenieurburo fur technische und naturwissenschaftliche Obersetzungen Dipl.-Obers. Dipl.-Dolm. Gerhard Herrera (BDO). Statthalterhofweg]1 (9).50858 Cologne Germany

Proof-reader:

Translator:

Dr. Ian Cole Aston University Birmingham United Kingdom

MA. MSc Stephen Day The Coach House. High Street Yalding. Kent ME 18 6HU United Kingdom

Translated from the German Edition ''Fahrzeuggetriebe'' ISBN3-540-65903-X Springer-Verlag Berlin Heidelberg NewYork Library Congress Cataloging-in-Publication Data applied for Die Deutsche Bibliothek - CIP-Einheitsaufnahme Lechner, Gisbert: Automotive transmissions: fundamentals, selection, design and application / Gisbert Lechner; Harald Naunheimer. In collab. with Joachim Ryborz. -Berlin; Heidelberg; New York; Barcelona; Hong Kong; London; Milan; Paris; Singapore; Tokyo: Springer, 1999 Einheitssacht.: Fahrzeuggetriebe <eng!.> ISBN 3-540-65903-X This work is subject to copyright. All rights are reserved, whether the whole or part of the material is concerned. specifically the rights of translation. reprinting. reuse of illustrations. recitation. broadcasting, reproduction on microfilm or in other ways, and storage in data banks. Duplication ofthis publication or parts thereofis permitted only under the provisions ofthe German Copyright Law of September 9. 1965. in its current version. and permission for use must always be obtained from Springer-Verlag. Violations are liable for prosecution act under German Copyright Law. Springer-Verlag Berlin Heidelberg New York a member of BertelsmannSpringer Science+Business Media GmbH Springer-Verlag Berlin Heidelberg 1999 Printed in Germany

@

The use of general descriptive names. registered names. trademarks. etc. in this publication does not imply, even in the absence ofa specific statement. that such names are exempt from the relevant protective laws and regulations and therefore free for general use. Typesetting: camera-ready by authors Cover design: medio GmbH. Berlin 62/3111-54321 - Printed on acid-free paper

Preface to the English Edition

"Automotive Transmissions" was first published in German in May 1994. It was so well received that we decided to publish the book in English, especially in view of the trend to market globalisation. This book gives a full account of the development process for automotive transmissions. It seeks to impart lines of reasoning, demonstrate approaches, and provide comprehensive data for the practical task of developing automotive transmissions. Much of the content is concerned with aspects of technology and production in the field of automotive transmissions that are of general validity, and hence of enduring relevance. The dynamics of the automotive transmission market since 1994 is reflected in numerous new types of transmission. Principal factors include the increasing use of electronics, light-weight construction, and the automation of manual gearboxes. Chapter 12 considers numerous production designs. to illustrate the theoretical principles· expounded in the earlier chapters, considering the main types of transmission and examining important detail solutions incorporated in specific mechanisms. Today's current design engineering is no longer state of the art tomorrow, with the next change of model. The designs presented here therefore claim to represent the different types of transmission design considered, rather than the latest production technology. Certain aspects of the book relate to the situation in Germany, particularly as regards transport systems and the economic significance of motor vehicles. This English language edition could not have come to fruition without the assistance of many contributors. We are particularly indebted to Dipl-Ing Joachim Ryborz as the manager and co-ordinator of the project, and to his assistants at the Institute of Machine Components (lMA), University of Stuttgart. We would also like to thank Stephen Day of Ubersetzungsbiiro Herrera for his professional translation of this book, and Dr Ian Cole for proof-reading the final text. We also wish to gratefully acknowledge the consistent financial support of the following companies: Audi AG, Robert Bosch GmbH, BMW AG, DaimlerChrysler AG, Ford Werke AG, GETRAG Getriebe- und Zahnradfabrik Hermann Hagenmeyer GmbH & Cie, Krupp Berco Bautechnik GmbH, Mannesmann Sachs AG, Adam Opel AG, Dr.-Ing. h.c. Porsche AG, Renk AG, Volkswagen AG, and ZF Friedrichshafen AG.

Stuttgart and Augsburg February 1999

Gisbert Lechner Harald Naunheimer

IJ - - - - - -..

I--------------------

""I........

Preface to the German Edition

It was in 1953 that H. Reichenbacher wrote the first book on motor vehicle transmission

engineering. At that time the German motor industry produced 490 581 vehicles including cars, vans, trucks, busses and tractor-trailer units. In 1992 production had reached 5.2 million. The technology at that time only required coverage of certain aspects, and Mr Reichenbacher's book accordingly restricted itself to basic types of gearbox, gear step selection, gear-sets with fixed axles, epicyclic systems, Fottinger couplings and hydrodynamic transmissions. Automotive engineering and the technology of mechanism design have always been subject to evolution. The current state of the art is characterised by the following interrelations: Environment Traffic Vehicle Transmission. Questions such as economy, environment and ease of use are paramount. The utility of a transmission is characterised by its impact on the traction available, on fuel consumption and reliability, service life, noise levels and the user-friendliness of the vehicle. There are new techniques which now have to be taken into account, relating to development methodology, materials technology and notably strength calculation. Examples include serviceability calculations, the introduction of specific flank corrections, taking account of housing deformation, and the need for light-weight construction. Transmission design engineering bas been enriched by numerous variants. The manual two-stage countershaft transmission, preferred for longitudinal engines, and the single-stage countershaft transmission preferred for transverse engines now have many subvariants, e.g. automatic transmissions, continuously variable transmissions,torque converter clutch transmissions, twin clutch transmissions, and transmissions for all-wheel drive. The engine and transmission must increasingly be considered as one functional unit. The terms used are "power train matching" and "engine/transmission management". This can only be achieved by an integrated electronic management system covering the mechanical components in both engine and transmission. The technique of Systematic Engineering developed in the 1960's, and the increasing use of computers for design, simulation and engineering (CAD), are resulting in ever-reducing development cycles. This trend is reinforced by competitive pressures. Systematic product planning is another significant factor in this regard. It was therefore necessary to· create an entirely new structure for the present book "Automotive Transmissions". Modern developments have to be taken into account. The great diversity and range of issues in developing transmissions made it difficult to select the material for this completely new version of "Automotive Transmissions", especially within thp nrev::Iilin a cnnstf::lints. Nnt eVP.fV tonic cnnlrl he covered in clet::lil. Tn those

I

JI..-.

places whe;e there i; an established literatu;e, the authors have chosen to rely on it in the interests of brevity. '..__.,.....-------_--

_

The purpose of this book is to describe the development of motor vehicle transmissions as an ongoing part of the vehicle development system. Only by actively taking this interacuon into account is It possible to arnve at a fully viable transmission design. The aim is to highlight the basic interrelations between the drive unit, the vehicle. and the transmission on the one hand, and their functional features such as appropriate gear selection, correct gear step, traction profile, fuel consumption, service life and reliability on the other. Of course another major concern was torepresent the various engineering designs of modern vehicle transmissions in suitable design drawings. The book is addressed to all engineers and students of automotive engineering, but especially to practitioners and senior engineers working in the field of transmission development. It is intended as a reference work for all information of importance to transmission development, and is also intended as a guide to further literature in the field. Without the assistance of numerous people this book would not have been written. We would like to thank Dr Heidrun Schrapel, Mr Wolfgang Elser; Dr Ekkehard Krieg, Dr Winfried Richter, Mr Thomas Sparl, Mr Thilo Wagner, Dr Georg Weidner and Prof Lothar Winkler for researching and revising chapters. We also wish to acknowledge the contribution of numerous assistants and postgraduates for important work on specific aspects. We wish to thank Christine Habich for her professional editing. We would like to thank many employees and scientific assistants of the IMA (Institut fUr Maschinenelemente) for reviewing and checking various parts of the text. . Such a book cannot be published without current practical illustrations. The publishers wish to acknowledge their gratitude to numerous companies for making illustrations available: Audi AG, BMW AG, Eaton GmbH, Fichtel & Sachs AG, Ford Werke AG, GETRAG, Mercedes-Benz AG, Adam Opel AG, Dr.-Ing. h.c. Porsche AG, and Volkswagen AG. We are particularly indebted to ZF Friedrichshafen AG who have always been mostforthcoming in responding to our numerous requests for graphic material. We. are also indebted to Springer-Verlag for publishing this book. We would particularly·liketo thank Mr M Hofmann, whose faith in our project never wavered, and whose gentle but frrm persistence ensured that the book did indeed reach completion. Dr Merkle . then prepared the work for printing. We must also thank the publisher of the "Design Engineering Books" series, Prof Gerhard Pahl for his patience and advice. Our thanks especially to our families for their understanding and support.

Stuttgart May 1994

Gisbert Lechner Harald Naunheimer

Contents

Terms and Symbols 1

2

XVII

Introduction

1

1.1 Preface : 1.2 History of Vehicle Transmissions 1.2.1 Fundamental Innovations 1.2.2 Development of Vehicles and Drive Units 1.2.3 Stages in the Development of Vehicle Transmissions 1.2.4 Development of Gear-Tooth Systems and other Transmission Components 1.2.5 Development of Torque Converters and Clutches 1.2.6 Investigation of Phenomena: Transmission Losses and Efficiency 1.2.7 Overview

1 6 6 8 9 17 19 20 21

Overview of the Traffic - Vehicle - Transmission System

23

2.1

J

Fundamental Principles of Traffic and Vehicle Engineering 23 2.1.1 The Significance of Motor Vehicles in our Mobile World 24 2.1.2 Trends in Transport Engineering 28 2.1.3 Passenger and Goods Transport Systems 30 2.1.4 Alternative Transport Concepts 33 2.2 The Market and Development Situation for Vehicles, Gearboxes and Components 35 2.2.1 Market Situation and Production Figures 36 2.2.2 Development Situation 39 2.3 Basic Elements of Vehicle and Transmission Engineering 41 2.3.1 Systematic Classification of Vehicles and Vehicle Use 41 2.3.2 Why do Vehicles Need Gearboxes? 42 2.3.3 Main and Auxiliary Functions of Vehicle Transmissions, Requirements Profile ·..·..· ·.. 44 2.3.4 Interrelations: Direction of Rotation, Transmission Ratio, Torque 45 2.3.5 Road Profiles, Load Profiles, Typical Vehicle Use and Driver Types .; 49 2.4 Fundamental Performance Features of Vehicle Transmissions 4~ 2.4.1 Service Life and Reliability of Transmissions 50

--------,,--,---------~---------------i i

2.5 3

2.4.2 Centre Distance Characteristic Value 2.4.3 Gearbox Mass Characteristic Value 2.4.4 Gearbox Cost Characteristic Value 2.4.5 Transmission Noise 2.4.6 Gearbox Losses and Efficiency Transmission Design Trends

Power Requirement 3.1.1 Wheel Resistance 3.1.2 Adhesion, Dynamic Wheel Radius and Slip 3.1.3 Air Resistance : 3.1.4 Gradient Resistance 3.1.5 Acceleration Resistance 3.1.6 Total Driving Resistance 3.1.7 Efficiency Map ~ 3.2 Diversity of Prime Movers 3.2'.1 'Overview 3.2.2 Electric Drive 3.2.3 Hybrid Drive 3.3 Power Output, Combustion Engine Characteristic 3.3.1 Torque/Engine Speed Characteristic 3.3.2 Engine Spread, Throttle Map 3.3.3 Consumption Map

Power Conversion: Selecting the Ratios 4.1 Power Train 4.2 Speed Converter for Moving Off.. 4.3 Total Ratio and Overall Gear Ratio 4.3.1 Overall Gear Ratio , 4.3.2 Selecting the Largest Power-Train Ratio 4.3.3 Selecting the Smallest Power-Train Ratio 4.3.4 Final Ratio 4.4 Selecting the Intermediate Gears 4.4.1 Saw Profile Diagram 4.4.2 Geometrical Gear Steps 4.4.3 Progressive Gear Steps 4.5 Continuously Variable Transmissions

5

".

.

,

Mediating the Power Flow 3.1

4

51 S1

53 54 54 56 58

,

;

58 58 60 61 63 63 64 65 68 68 69 70 71 72 74 75 77 78 79 81 82 83 84 87 87 88 88 89 ; 90

Matching Engine and Transmission

92

5.1

94 95

Traction Diagram 5.1.1 DeriVing a Traction Diagram (Example)

XI

Contents

II

5.2

j

5.3

Ii

5.4 5.5

6

5.1.2 Engine Braking Force 5.1.3 Geared Transmission with Dry Clutch 5.1.4 Geared Transmission with Trilok Converter Vehicle Performance 5.2.1 Maximum Speed 5.2.2 Climbing Performance 5.2.3 Acceleration Performance Fuel Consumption 5.3.1 Calculating Fuel Consumption (Example) 5.3.2 Determining Fuel Consumption by Measurement.. 5.3.3 Reducing Fuel Consumption 5.304 Continuously Variable Transmissions EmIssIons Dynamic Behaviour of the Power Train, Comfort

Veh·ICI · · Syst ems: B· . Ies e T ransmlsslon aSlc D· eSlgn p. rmclp 6.1

6.2

6.3

6.4 6.5 6.6

6.7

·

97 98 98 101 102 103 103 104 104 106 107 107 108 109 111

Arrangement of the Transmission in the Vehicle 111 6.1.1 Passenger Cars III 6.1.2 Trucks and Buses 114 6.1.3 Four-Wheel Drive Passenger Cars 114 6.1.4 Transverse and Longitudinal Dynamics with All-Wheel Drive 119 Transmission Formats and Designs 120 6.2.1 Transmission Format 120 6.2.2 Transmission Design 121 Basic Gearbox Construction 123 6.3.1 Shifting with Power Interruption 124 6.3.2 Shifting without Power Interruption 124 6.3.3 Continuously Variable Transmissions without Power Interruption 125 Gear-Sets with Fixed Axles, Countershaft Transmissions and Epicyclic Gears 126 Fundamental Approaches for Part Functions, Evaluation 128 6.5.1 Reverse Gear ······ 129 Passenger Car Transmissions 130 6.6.1 Manual Passenger Car Transmissions 130 6.6.2 Semi-Automatic Manual Passenger Car Transmissions 133 6.6.3 Fully Automatic Passenger Car Transmissions 134 ;. 141 6.6.4 Continuously Variable Passenger Car Transmissions Commercial Vehicle Transmissions ,145 6.7.1 Single-Range Transmissions 146 6.7.2 Multi-Range Transmissions ·.·············· 147 154 6.7.3 Practical Design of Two- and Three-Range Transmissions 6.7.4 Semi-Automatic Manual Commercial Vehicle Transmissions 157 h.75 Fullv Automatic Commercial Vehicle Transmissions 158 6.7.6 Continuously Variable Transmissions for Commercial Vehicles lj~

I-1~-----T"lr------------------"'--

7

6.8 6.9

Transfer Gearboxes and Power Take-Offs Final Drives: Formats, Performance Limits, Transmission Ratios.......... 6.9.1 Final Drive Systems for Passenger Cars 6.9.2 Final Drive Systems for Commercial Vehicles 6.10 Differential Gears, Differential Locks and Locking Differentials 6.10.1 Principles of Differential Gears 6.10.2 The Need for Locking 6.10.3 The Interlock Value 6.1004 Alternatives to Self-Locking Differentials

160 1n2 163 164 167 168 169 170 171 .

Design of Gearwheel Transmissions for Vehicles

173

7.1

7.2 7.3 7A

7.5

8

Gearwheel Performance Limits 7.1.1 Causes and Types of Damage 7.1.2 Calculating the "Tooth Failure" Performance Limit. 7.1.3 Calculating the "Pitting" Performance Limit 7.1.4 Calculating the "Gear Scuffing" Performance Limit Estimating Centre Distance Estimating Face Widths Operational Integrity and Service Life 7.4.1 The Wohler Curve 7.4.2 Load Profile and Enumeration 704.3 Damage Accumulation Hypothesis Developing Low-Noise Transmissions 7.5.1 Transmission Noise and Its Causes 7.5.2 How Noise Reaches the Ear 7.5.3 Assessment Criteria 7.5.4 Countermeasures

Specification and Design of Shafts . 8.1

Typical Problems in Vehicle Transmissions 8.1.1 Configuration of Shafts in Vehicle Transmissions 8.1.2 Designing for Stress and Strength 8.1.3 Deflection 8.104 Vibration Problems 8.2 General Design Guidelines 8.3 Transmission Drive-Shaft Strength Design 8.3.1 Loading 8.3.2 Bearing Reactions 8.3.3 Spatial Beam Deflection 8.304 Power and Torque Profiles 8.3.5 Critical Cross-Section 8.3.6 Stresses 8.3.7 Preliminary Specification of the Shaft Diameter

173 174 178 178 180 180 183 184 185 187 189 195 195 199 199 202 204 204 204 204 206 206 207 208 208 211 211 212 214 215 218

~I

..•.......

....

:

:

I-l

804 8.5

9

218 219 220 220 221

Gear Shifting Mechanisms, Layout and Design of Synchronisers

224

9.1

226

9.2

9.3 9A

9.5 9.6 9.7

10

8.3.8 Designing for Fatigue Strength 8.3.9 Designing for Operational Integrity 8.3.10 Normal Shaft Materials Calculating Deformation Flow Chart for Designing Transmission Shafts

Systematic Classification of Shifting Elements 9.1.1 Shifting Elements for Geared Transmissions with Power Interruption 9.1.2 Shifting Elements for Geared Transmissions without Power Interruption 9.1.3 Parking Lock Synchroniser Functional Requirements 9.2.1 Changing Gear 9.2.2 Main Functions and Ancillary Functions 9.2.3 Speed Synchronisation with Slipping Clutch 9.204 Synchroniser Dimensions The Synchronising Process 9.3.1 Ease of Use Design of Synchronisers 904.1 Synchroniser Performance Limits 904.2 Basis for Design Calculation 904.3 Practical Design for Acceptable Thermal Stress 904.4 Designing Locking Toothing for Locking Effect The Tribological System 9.5.1 Materials Engineering Designs 9.6.1 Detail Questions Alternative Transmission Synchronisers

226 229 230 231 233 234 234 236 237 239 241 241 244 245 249 253 253 254 258 259

Hydrodynamic Clutches and Torque Converters 10.1 Principles 10.2 Hydrodynamic Clutches and their Characteristic Curves 10.3 Torque Converters and their Characteristic Curves 10.3.1 The Trilok Converter 1004 Engine and Torque Converter Working Together 1004.1 Torque Converter Test Diagram, Interaction of Engine and Trilok Converter 10.5 Practical Design of Torque Converters 10.6 Engineering Designs 10.7 Design Principles for Increasing Efficiency 10.7.] Torqpe Co;!verter LoC'k-TJp rlntrl-] . _..... .. . . 10.7.2 Power Split Transmission

261 ;.. 262 265 266 267 268

~

270 272 272 274 274 · 275

11

Notes on the Design·and Configuration of Further Vehicle Transmission Design Elements

279

11.1 Bearings 11.1.1 Selecting Bearings 11.1.2 Bearing Design 11.1.3 Design of Roller Bearings 11.2 Lubrication of Gearboxes, Gearbox Lubricants 11.2.1 Bearing Lubrication 11.2.2 Principles of Lubricating Gearwheel Mechanisms 11.2;3 Selecting the Lubricant.. 11.2.4 Selecting Lubricant Characteristics 11.2.5 Lifetime Lubrication in Vehicle Gearboxes 11.2.6 Testing the Scuffing Resistance of Gearbox Lubricants 11.3 Gearbox Housing : 11.3.1 Gearbox Housing Design 11.3.2 Venting Gearboxes 11.4 Gearbox Sealing 11.4.1 Seals for Static Components 11.4.2 Seals for Rotating Components 11.4.3 Seals for Reciprocating Round Components 11.4.4 Practical Examples 11.5 Vehicle Continuous Service Brakes 11.5.1 Definitions 11.5.2 Engine Braking Systems 11.5.3 Retarders 11.5.4 Actuation and Use

279 280 280 284 286 287 287 290 290 293 294 295 295 297 301 301 304 305 306 307 308 309 309 313

12 Typical Designs of Vehicle Transmissions 12.1 Manual Gearboxes 12.1.1 Manual Passenger Car Gearboxes 12.1.2 Manual Commercial Vehicle Gearboxes 12.2 Semi-Automatic Manual Gearboxes 12.2.1 Semi-Automatic Manual Passenger Car Gearboxes 12.2.2 Semi-Automatic Manual Commercial Vehicle Gearboxes 12.3 Fully Automatic Gearboxes 12.3.1 Fully Automatic Passenger Car Gearboxes 12.3.2 Fully Automatic Commercial Vehicle Gearboxes 12.4 Further Examples 12.5 Final Drives 12.5.1 Typical Designs, Passenger Cars 12.5.2 Typical Designs, Commercial Vehicles 12.6 Differential Gears, Locking Differentials · · 12.7 Four-Wheel Drives, Transfer Gearboxes ··

314 315 315 322 326 326 : 326 328 329 332 333 340 340 343 ·.. 346 · · 352

I-1,

Contents

13

14

15

16

XV

Engine and Transmission Management, Electronics ~ and Information Networking

359

13.1 Overview of Electronic Systems in Current Use 13.2 Engine Management 13.3 Transmission Control. 13.3.1 Automatic Master/Gearshifting Clutch 13.3.2 Semi-Automatic Manual Transmissions, Automatic Gear Selection. 13.3.3 Fully Automatic Transmissions, Adaptive Gearshift Strategy 13.3.4 Continuously Variable Transmissions 13.4 Electronically Controlled Braking and Traction Systems 13.5 Safety Concepts

359 361 361 361 362 362 364 364 364

Overview of the Development Process, Product Planning and Systematic Engineering Design

365

14.1 14.2 14.3 14.4 14.5

366 368 371 373 380

Product Life Cycles Product Planning The Development Process Systematic Engineering Linking Development and Production

Computer-Aided Transmission Development, Driving Simulation

381

15.1 Driving Simulation : 15.1.1 Extraneous Factors 15.1.2 Route Data Set, Route Data Acquisition 15.2 Driving Simulation Programs 15.2.1 Classification 15.2.2 Modular Construction 15.3 Applications of Driving Simulation

383 384 385 386 386 387 388

Reliability and Testing of Vehicle Transmissions

391

16.1 Principles of Reliability Theory 16.1.1 Definition of Reliability 16.1.2 Statistical Description and Representation of the Failure Behaviour of Components 16.1.3 Mathematical Description of the Failure Behaviour Using the Weibull Distribution 16.1.4 Reliability with Systems 16.1.5 Availability of Systems 16.2 Reliability Analysis of Vehicle Transmissions 16."'.. 1 Sy~tE'1T1 An;'11y"i~ 16.2.2 Qualitative Reliability Analysis

392 392

-_

392 395 400 402 403 403 406

16.2.3 Quantitative Reliability Analysis ;.~ 16.3 Testing to Ensure Reliability 1 O.j.1 Classitylllg Vehicle Transmission Test Programs '" 16.3.2 Test Stands for the Test Programs

::

408 411 412 415

References Index of Companiesffransmissions

419

;

439

Index of Names

441

Subject Index

442

Terms and Symbols

A formula you cannot derive is a corpse in the brain Ie. WEBER!

Physical variables· are related by mathematical formulae. These can be expressed in two different ways:

o o

quantity equations, unit equations.

Quantity Equations Quantity equations are independent of the unit used, and are of fundamental application. Every symbol represents a physical quantity, which can have different values: Value of the quantity = numerical value x unit. Example: Power P is generally defined by the formula P=Tw,

(1)

where T stands for torque and w stands for angular velocity.

Unit Equations If an equation recurs frequently or if it contains constants and material values, it is convenient to combine the units, in which case they are no longer freely selectable. In unit equations the symbols incorporate only the numerical value of a variable. The units in unit equations must therefore be precisely prescribed.

Example: In order to calculate the effective power P in kW at a given rotational speed n in lImin, the above equation (1) becomes the unit equation

P

=.!...!:- . 9550

(2)

Terms and Symbols (Only those which occur frequently; otherwise see text) A

AR BIO Bx

C CC CG CGH CGL CGmain CGR D E F

Fa

FB FH FL Fn

FQ FR Fr Fs

FS t Ft

Fu Fz F(t)

GR J

KG L

Mb Mt My N P PA Pe Pm PZ,B

Q R Ra

Surface area, transverse couple surface area =projection of vehicle front area Synchroniser friction contact System service life for a failure probability of 10% System service life for a failure probability of x% Pitch point, dynamic contact figure, constant Torque converter lock-up clutch Constant gear Front-mounted splitter constant high Front-mounted splitter constant low Main gear unit constant Range constant Diameter Modulus of elasticity Force Acceleration resistance, axial force Braking force Slope negative lift force Air resistance, bearing force Normal force Transversal force Wheel resistance Radial force Lateral force Gradient resistance Tangential force Circumferential force Traction Distribution function, failure probability Wheel load Mass moment of inertia Gear characteristic value Service life Bending moment Torsional moment Reference moment Fracture cycles Power Friction power related to area (synchroniser) Effective power at engine output Average friction work during synchroniser slipping times Demand power at wheel Lateral force, volume flow Reaction force Average peak-to-valley height

J

Terms and Symbols

Re Rm R(t) S SB SH SL ST

T TB TL TR TC TCC U

VH W WA Wb Wt a

b bo be bs bs C Cs Cu

ew Cy

d e

f fR fit)

g hi 1

lA lE lO lO, tot leo

lM IN

lS iy 1

k key)

XIX

Yield point Tensile strength Survivability, reliability Safety factor, locking safety with synchronisers, slip, interlock value Brake slip Rear-mounted splitter unit high Rear~mounted splitter unit low Drive slip Torque, characteristic service life Acceleration torque (synchroniser), locking torque (differential) Load torque Friction torque (clutch, synchroniser), reactor torque (torque converter) Torque converter Torque converter clutch Revolutions Total displacement Moment of resistance, work, usable work, friction work (Specific) friction work per unit area Moment of resistance against deflection Moment of resistance against torsion Acceleration, axle base Form parameter, failure gradient, overall length, width, fuel consumption Size factor Specific fuel consumption Surface factor Fuel consumption per unit of distance Constant, rigidity, absolute speed Tooth spring rigidity Circumferential component of absolute speed Drag coefficient Average value of tooth spring rigidity over time Diameter Eccentricity Deflection Coefficient of rolling resistance Density function Acceleration due to gravity Load cycle Ratio Power-train ratio (from engine to wheels) Final ratio Gear ratio Overall gear ratio, ratio spread Constant gear ratio Centre gear ratio Hub gear ratio Moving-off element ratio Variator ratio Number of friction contacts Wohler curve equation exponent Characteristic value of a torque converter

II

m mF

n nM p q

q' r rdyn S

SFn

t

to tm tR

ts u V

VF Vth

Vw w

x,y, Z Z Zj

a aOK aK ak

an aSt

ao j3 j3 K

LI LIS LlV 8 8a 86

TJ

9A A(t)

P Po PG PH

Gear modulus, mass, linear scale (converter) Vehicle mass Rotational speed, quantity, stress reversals, numher of hearim7 s Engine speed , . ~ Contact pressure, pressure, number of gear pairs Gradient, surface load Gradient in % Radius, degree of redundancy Dynamic tyre radius Travel, gearshift sleeve travel, fin pitch Root thickness chord Statistical variable, time Time without failure Mean of Weibull distribution Slipping time, friction time Shifting time Gear ratio, circumferential speed Speed, flow rate Vehicle speed Theoretical speed where A =0 Wind speed Work input, relative wind speed Co-ordinates Number of speeds, number of teeth, number of load cycles Number of teeth gear i

Meshing angle, taper angle of a taper synchroniser, viscosity pressure coefficient Throttle valve angle Force meshing angle relative to tip edge Statistical form factor Normal meshing angle Gradient angle Strain ratio Helix angle at pitch circle, aperture angle of claws Dynamic beam stress rate Interval, difference Wear path Wear Total contact Transverse contact ratio Overlap ratio Efficiency, dynamic viscosity Temperature , Performance coefficient (converter, retarder), drive slip, rotational inertia coefficient Failure rate Torque conversion, coefficient of friction Stall torque ratio Coefficient of sliding friction Coefficient of bonding friction

XXI Speed ratio, kinematic viscosity Density, angle of friction of the claws Direct stress Bending stress Fatigue strength Hertzian stresses Reference stress Torsional stress, torque increase with combustion engine Gear step, bending angle Basic step with progressive stepping Progression factor with progressive stepping Gear step with geometrical stepping Angular velocity

Subscripts

o

U

Nominal or initial state Pinion (= small gear), input Wheel (= large gear), output At point 1,2, 3, ... Offer, related to area, drive shaft, power train, moving off Demand, brake Clutch Constant gear Countershaft Duration, fatigue-resistant, deficit, opening, direct drive End Excess Vehicle, root, free-wheel Gearbox Adhesion, main, main gearbox, main shaft wheel, ring gear, high (= fast) Input shaft Air, load, low (= slow) At bearing point, at bearing point 1, 2 Engine, motor, model Main shaft Rear-mounted range-change unit Output shaft Pump, pump wheel, planetary step Pump test Transverse Reverse gear, roll, slip, friction, wheel, range-change unit, reactor Roll Status, system, splitter unit Pulsating (strength) Gradient Turbine wheel, drive Torque converter Clrcumterence

V

F_ront,..-m_o_u_n_te_d_s_p_li_tt_e_r_u_ni_t,,_v_ar_._ia_t_or_,_lo....S_s_,_tr_ia_I

1 2 1, 2, 3, ... A B C CG CS D E Ex F G H IS

L L,Ll,L2

·M MS N OS P PV

Q

.. J ! :_.

R Roll S Sch St

T TC y,

_

w z

Reversing (strength) Traction, tensile load, intermediate gear, tooth, opening

a abs b dyn e fric front fuel rear

Acceleration, axial, values at tip circle, tip of gear, outlet, external Absolute Bending Dynamic Effective, inlet Friction Front Fuel Rear Internal, control variable i = 1, 2, 3, ... n Ideal Input At point i, j Control variable Control variable Beam stress Mean, number of stress classes Main Maximum Minimum n-th speed, nominal, nominal operating point Oil Output Permissible Radial Reduced Reference Relative Resultant Specific Static Torsion, time Theoretical Total Twist Reversing, pitch circle In x, y, z direction, around x, y, z axis. Highest speed, number of speeds

1

id In 1, J J k kt m maIn max mIn n

oil out perm r

red ref reI res spec stat t

th tot twist w x,y,z z

-1,I

1 Introduction

Every vehicle needs a transmission!

1.1 Preface All forms of motorised transport, including vessels and aircraft, need transmissions to . convert torque and rotation (Figure 1.1). There are distinctions between transmissions according to their function and intended use, for example selector gearboxes, steering boxes and power take-offs. This book deals only with road vehicle transmissions, or transmissions for vehicles for combined road and off-road use (outlined in bold in Figure 1.1). Figure 1.2 gives an overview of transmission types in general current use. Further details are given in Chapter 6 "Vehicle Transmissions Systems: Basic Design Principles". The function of a vehicle transmission is to adapt the traction available from the drive unit to suit the vehicle, the surface, the driver and the environment. The main parameters are technical and economic competitiveness. The transmission has a decisive effect on the reliability, fuel consumption, ease of use, road safety and transportation performance of passenger cars and commercial vehicles (Figure 1.3).

Vehicles

I Vessels

Land vehicles •...

Aircraft

"

......•....

I Road-going vehicles

Construction vehicles

I

I

vehicles

Tracked and special-pu rpose vehicles

~gricultural

I Rail vehicles

I Construction vehicles on+off-road use

Exclusively on-road use

I ,:.",-;.- .>':":,"::;;;.;:

>,j:t::}~

.~~

~earb

I

I

Power take-offs

Final drives

I Transfer boxes Differential gears Differential locks

I Steering boxes

Figure 1.1. Definition of the term "automotive transmission" for the purposes of this book

_

Transmission types z-speed-transmissions: (geared transmissions with z speeds) .c

enC

Q)O

E'w ,en c'E (ljen

wC

C(lj o~

()

.cc

.c C

eno Q).E~ 0'-

o.Q -en

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I

,~

Oc

C

o

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E =E :::l en

en

~ C

C(lj >, ...

Q..

c

- -

I-~

en-

With power interruption Moving off and disengagement with foot-operated clutch Manual gearshift

Continuously variable transmissions

~

_.-en m -en

cl::"

(lj'~

'0 en 0) '0 'E- C ... 0 o cen.I- (lj.(3

......

-::=..

c o 0 ' en ._ en·- ... en E o E ... en oQ) en ~c Wc

o

C

.- 0 (lj

_.-

- -

I~

Without power interruption (powershift)

Automatic continuously variable moving off

Semiautomatic gearshift

Automatic gearshift

Automatic torque and speed conversion

Figure 1.2. Systematic classification of vehicle transmissions

Economy

Reliability Service life

Fuel consumption

Transport capacity

Road safety Ease of operation

~

Figure 1.3. Effect of the transmission on key features of the vehicle

~

I-.l

1.1

Preface

3

100% Partly developed technologies

1 (J)

Q)

ca

Highly developed technologies

Monorail, computer

Consumer goods

Vehicles, gearboxes, machine tools, antifriction bearings

(J)

c: 0

+: 0

c:

:::J

Refrigerators, binoculars, detergents, bicycles

cD

(J)

::J

Large...

Moderate increase

Small...

Development, production and marketing resources required -

Figure 1.4. Achievable increase in utility of a product by additional development effort

Vehicle transmissions are technically and technologically highly mature mass-produced products. They are categorised as highly developed technologies (Figure 1.4). One notable feature is the specific power handling capacity P spec in kW/kg of vehicle transmissions, which is more than twice that of industrial transmissions (Table 1.1), despite the fact that vehicle transmissions have more speeds. But industrial transmissions have to be designed for longer service life. There are unlikely to be any further fundamental innovations in vehicle transmission technology. There is more likely to be a process of gradual evolution. The main trends are system thinking embracing the factors Environment ¢:::> Traffic ¢:::> Vehicle ¢:::> Transmission, and greater use of electronics for control and monitoring processes. This defines the superordinate development goals for vehicle transmissions (Figure 1.5). Their development has to be fast and market-orientated. There has to be flexibility in adapting to customer preferences, especially in the case of commercial vehicles. Legal requirements also have to be taken into account, such as kW/t constraints and the maximum permissible noise level. Table 1.1. Comparison of a vehicle transmission (commercial vehicle) with an industrial transmission Transmission

Number of stages/ gears

Ratio

Power P(kW)

Input torque T1 (Nm)

Volume (m 3 ) Mass (kg)

Specific power Pspec (kW/kg)

0.2823

0.479

680

100%

0.159

1.06

'l')r

221 C;~

Industrial use Two-stage 12.5

330

2100

1 Gear Automotive use

Two-stage or 3-stage 13.8 16 gears in 1st

~- I

i

r0\i0iSe

gear

[~'

356

1700 vv...J

(

Fast, flexible

(

I

Satisfy legal

J

l.. .__

development

re_q_u_i_re_m_e_n_ts_ _.....

~ - - - - - - - - - -..... Superordinate development goals

Take account of technical developments

I Optimal conversion of the engine power available: good economy and acceleration

High reliability, adequate service life, low noise and environmentally sound

Figure 1.5. Superordinate development goals for vehicle transmissions The main goal when developing a vehicle transmission is to convert the power from the engine into vehicle traction as efficiently as possible, over a wide range of road speeds. This has to be done ensuring a good compromise between the number of speeds, climbing performance, acceleration and fuel consumption of the vehicle. Technical and technological advances have to be taken into account, as do operational reliability and adequate service life. It is also essential to have regard for environmental and social considerations. Vehicle transmissions always have to be developed within the framework of planning horizons fm: new vehicles (Figure 1.6). Transmissions have to be developed or adapted in parallel with the development phase of a vehicle. This also involves preparing and introducing new production technologies for mass production.

Development phase

o

2

4

6

8

10

12

14

16

Years

20

Figure 1.6. Time dimensions and planning horizons in the automotive industry as per [1.1]

1.1

Preface

5 This. boo~ sets out to pr~sent .the .development process for vehicle transmissions in its totalIty (FIgure 1.7). The mtentIOn IS to put the development of vehicle transmissions into its broader ~ontext. It is always necessary to determine the overall system within which the product IS to be used, apart from just the product. A system overview is essential and ' is presented in Chapter 2. The critical constraints for the vehicle transmission designer are the vehicle the engine and the operating environment. Basic knowledge of these factors is essenti;l for meaningful development. Chapter 3 highlights the interaction between the power required and the power available. The first specific development task involving the gearbox is then selecting the range of ratios, or "overall gear ratio", to be covered. The functioning of vehicle and transmission as a system can then be assessed, and decisions made as to the number of speeds z, the ratio of the individual speeds, and the resultant gear stages. Taking into account the operating environment, the designer then has to decide whether the vehicle has adequate acceleration, the required climbing performance, and the top speed U max stipulated in the specification. This also determines whether the transmission enables efficient motoring, especially in terms of fuel consumption. This issue is examined in detail in Chapters 4 and 5. Creative design remains essential, and is now supported by systematic engineering design. This involves carrying out a functional analysis at the concept stage. Solutions must be found and assessed for the various individual functions, and then combined to form an overall solution. The necessary knowledge of vehicle transmission systems is given in Chapter 6. Chapters 7 to 11 examine the design and construction of the main components of a transmission: gearwheels, shafts, bearings, synchronisers, and hydrodynamic clutches and torque converters. The sophisticated techniques now in use, such as the finite element method (FEM) and gearwheel calculation to German standard DIN 3990, are not examined in detail. An attempt is made to describe the fundamentals of design methodology and engineering processes.

r--

---:-:--:-1

Engine, vehicle and environment as parameters

3

Overall gear ratio, ratios, traction diagram Product planning and systematic . design Electronics and information networking

I I I __ lTIJ---"""""

Examples of vehicle transmissions in use, model designs :':1'2':

Vehicle transmission systems Design of main components

Figure 1.7. Map of the processes involved in developing vehicle transmissions, overview of chapters

Chapter 12 examines the construction of various types of transmission, with a detailed review of numerous existing designs and key design elements. No treatment of the subject would be complete without considering the increasing use of electronics. Developments are leading to integrated engine/transmission management (Chapter 13). Electronics is also being applied to assist the shifting process in conventional manual transmissions and in automatic transmissions. Important development tools for designing vehicle transmissions are dealt with in the latterpart of this book. Product planning, project planning and systematic engineering are examined in Chapter 14. Chapter 15 considers the use of computer-aided design (CAD) and driving simulation for optimising transmission design. Product reliability is becoming increasingly important. The customer is principally concerned with the reliability and service life of the system as a whole. Extensive testing on suitable test rigs is essential for determining service life and reliability. This topic is dealt with in Chapter 16. A special attempt has been made in this bookto show the user how to proceed, and to provide comprehensive data for practical transmission development work. As DUDECK has stated, "One ofthe tasks of engineering science is to refine complicated models to the point of simplicity". This book strives towards that aim.

1.2 History of Vehicle Transmissions

Knowledge of the past and of the state of the Earth affords the human spirit delight and sustenance ILEONARDO DA VINCII

Applying the lessons of the past! Development engineers and designers should have a grasp of the historical development of their products. They can then assess what yet remains to be achieved, and the level of technology represented by current product development. Such knowledge complements systematic design (see Chapter 14).

1.2.1 Fundamental Innovations Fundamental innovations are discoveries, inventions and new developments without which the existing product could not have been developed. These seminal innovations inform subsequent discoveries, inventions, new developments and designs, leading to the creation of new products (Figure 1.8). The aim of such development processes is to investigate and research certainphenomena to ensure reliable operation of the product. Table 1.2 is an attempt to trace the development of basic mechanical engineering innovations which lead to motor vehicles and thus to vehicle transmissions. I

,..,j ,

I

1.2

History of Vehicle Transmissions

7

Basic innovations for vehicles and vehicle transmissions examples see Table 1.2

Inventions, new developments and research work derived from basic innovations

, ~

Development steps with vehicles and vehicle drives: examples see Table 1.3

~

Development steps with vehicle transmissions: examples see Table 1.4

~ if}'

Development steps of toothing and other gearbox components: examples see Table 1.5

~

Development steps with hydrodynamic converters and clutches: see Table 1.6

~

Investigation of phenomena: examples see Table 1.7 "Research into transmission loss phenomena"

Figure 1.8. Products are developed on the basis of fundamental innovations!

Table 1.2. Examples of seminal innovations for vehicles and vehicle transmissions

4000

BC 2500

BC 20001000

BC 500

BC 200

BC 1754

1,

1769 1/ ()q

Mesopotamian vase with apicture of a cart Wheels made of two semicircular wooden discs, presumably with leather tyres Spur gears with pin wheel gear as drive element for water scoops (Sakie, Figure 1.10), worm gears for cotton gins Greek scholars discover the principles of mechanics Lever, crank, roller, wheel, hoist, worm gear and gearwheel are in use Euler's law of gears for gearwheels, involute toothing Watt Patent for steam engine Walt Gearbox WIth constant-mesh g g e_n__a_e_m_,_en_t

.

1829 1877 1885 1897 1905 1907 1923 1925

Stephenson Rail vehicle, steam locomotive Otto Patent for 4-stroke gas engine with compression Benz Three-wheeler with internal combustion engine Bosch Magneto electric ignition Fottinger Hydrodynamic torque converter Ford Mass production of model T; the passenger car becomes a massproduced item Bosch Injection pump Rieseler Automatic passenger car transmission with torque converter and planetary gear-set

_

1.2.2 Development of Vehicles and Drive Units fhe Hiea of equipping lhe engine with a gear unil fur adapling" l"t'l-' "'"1 ." " . ' ~ torque to the pow t . " 01 gJlle , 'pew ano rs of the automobile yet before the official date of birth to convert the reciprocating movement of th . t em: y years of the engme was the need . . e pIS ons mto rotary movement. One solution is shown in Fi Ufe I 9 The. linked to that all e~gi~es. hlstoncal development of the transmission is thus closely

i~°l~~~t=~:~:r~~~e:o:::~~

0;

Figure 1.9. Conversion of reciprocating movement into rotary movement. Twincylinder power unit with opposed pistons in steam passenger car (CUGNOT, 1725 to 1804)

Table 1.3. Chronological development of vehicles and drive units

5000- First technical inventions known: 500 Be wheel, cart, gearwheel 1500 Diirer Sketch of a self-propelled vehicle 1690 Papin Designs an atmospheric steam engine with cylinder and pistons 1769 Cugnot Steam vehicle with rectifier transmission 1784 Watt Double-acting steam engine with rotary movement and flywheel 1800 1801 1801 1814 1817 1832

Trevithick Patent for high-pressure steam engine Trevithick Use of steam vehicle to carry passengers Artamo Metal bicycle with now pedal cranks Stephenson First stearn locomotive Drais Steerable road wheel Pixii Rotating alternating current generator

1845

Thompson Invention of the pneu-

matic tyre Lenoir Double-acting gas piston engine 1866 Siemens Discovery of the dynamoelectric principle and design of an operational dynamo 1877 Otto Patent for four-stroke gas engine with compression 1884 Parsons Patent for steam turbine 1885 Benz Three-wheeler with com-

1862

bustion engine Daimler Motorcycle DaimlerlMaybach Four-wheel motor car 1888 Dunlop Pneumatic rubber tyre 1889 Maybach-Daimler Steel wbeeled passenger car with open 2-Speed transmission 1897 Bosch Controlled electric magneto ignition

1885 1886

1.2

History o/Vehicle Transmissions

9

Diesel Diesel engine; heavy fuel engine with compression ignition Wright brothers Powered flight Ford Introduction of mass production line Gregoire Constant-velocity joint. The Tracta joint opens the door to mass-produced front-wheel drive

1897 1903 1907 1926

1934 1970 1980 1990

Porsche Project draft of the Volkswagen Thyssen Henschel Transrapid maglev monorail France TOV high-speed trains Bundesbahn ICE high-speed trains

1.2.3 Stages in the Development of Vehicle Transmissions Gear systems were undoubtedly used more than 1000 years ago to enhance the effectiveness of human and animal effort. Like the bullock gear systems, that are still in use in Egypt today, the two mating parts interlock by means of wooden pins or teeth (Figure 1.10). The first drawings of gear systems are from the Middle Ages. Muscle power was used in the absence of mechanical power. The human "machines" had to do the hard work. This was the origin of the first "vehicle transmissions". In Albrecht Diir.er's etching of a "muscle-powered vehicle" around 1500, the limited human power stroke is converted into propulsive force by means of a thrust crank, a bevel gear and a spur gear stage.

Figure 1.10. An early gear system! Egyptian water scoop (Sakie) in Luxor, approximately 2000 to 1000 Be Table 1.4 gives examples of the important stages in the development of vehicle transmissions. It shows that all the main elements and design principles for vehicle transmissions had been developed by 1925. Developments since that date have been aimed at improving service life and performance, and/or reducing weight and noise, and improving ease of use. Four lines of development can be distinguished (FIgure 1.11):

J ".-

o o o o

mechanical z-speed geared transmission, semi-automatic or fully automatic z-speed geared transmission, conventional hydrodynamic/mechanical geared automatic transmission, mpchani~21. h;rrlr(vlymHllie

tr_a_n_s_missl""'io_n_.

Tl

h;'or()C't'lti(' nr (? lp r: tr ;c

('0~,t;t;.~l()nsJ:T '.T",~j2b!c

- - - - - - - - - - - - -- - - - - - -

I b)

g)

h lID

h) C2

i)

Fi£ure 1.11. Development sequence of passenger car and commercial vehicle transmissi~. aJ Transmission with sliding gear engagement; b) Transmission with constantmesh et1gagement; c) Synchromesh gearbox; d) Torque converter clutch gearbox, semiauromaltic: converter + gearshifting clutch + synchromesh gearbox; e) Transmission with multipl.ate clutch shift; 0 Transmission with converter and rear-mounted powershiftable count.en>haft transmission; g) Hydro-planetary transmission; h) Conventional fully automatic transmission; i) Hydrostatic continuously variable transmission with power split, fully automatic; k) Mechanical continuously variable pulley transmission

Table lA. Examples of important development steps in vehicle transmissions

1784

:

Watt stipulates that. steam engines require additional ratios for roadgoing vehicles. Watt patents variable-speed gearbox with dog clutch engagement and constant mesh of gearwheels (Figure 1.12)

I

1821 1827 1834

Griffith 2-speed transmission with sliding gears (Figure 1.12) Pequeur First differential in a roadgoing vehicle (Figure 1.12) Bodmer Planetary transmission with stallable ring gear body using brake band

:, "~.~,,~.,i,,;.,

1.2

History of Vehicle Transmissions

11

'"":1

11

I i

I

1849 1879

1885 1886 1889 1890 1899 1899

I

II I

1899 1899 1900

I

1900 ,~

-j -\

!

,

-

J

1900 1900 1905 1906 1906

1907

Napier/Anderson 2-speed belt transmission (Figure 1.12) Selden Patent enclosed sliding gear transmission with reverse gear and clutch (Figure 1.12) Marcus Cone clutch for motor vehicles Benz Belt-driven bevel gear differential (Figure 1.12) Maybach-Daimler 4-speed transmission with sliding gears" (Figure 1.13) Peugeot Complete power train with sliding gear drive (Figure 1.13) Buchet Continuously variable belt transmission with axially adjustable taper discs Krauser/Schmidt Continuously variable friction gear with taper discs Darracq - Leon - Bollee 5-stage variable-speed belt "transmission gearbox" Oliverson - Killingsbeck Continuously variable belt transmission with axially adjustable taper discs Reeves - Pulley Continuously variable belt expanding pulley transmission with thrust links and axially adjustable taper discs Leo 3-speed transmission with face dog clutch engagement, integral differential and chain drive reverse gear Lang 3-speed geared transmission with constant-mesh wheels and draw key shifting Diamant Speed Gear Company Helical gear transmission PUtler Hydraulic drive system with hydro pump and hydro motor Renault Pneumatic transmission with piston compressor and piston engme Didier Two-stage planetary gear transmission with shifting using brake band and clutch of the planetary gear via friction plate face clutch Renault Hydrostatic transmission T,T,':~h 2;:i~! ristC'~~~lI~r a:~d ~xi::,.J

piston motor

1907 1915

1925 1925

1926

1928

1928

1929

1931 1932

1934

1939

1939

10t1Q ,> ; J~ _

~'

Ford Mass production of the model T with 2-speed planetary gear ZF-Soden transmission 4-speed all constant-mesh transmission with constant-mesh gearwheels with preselector shifting and with synchronising aids ZF Commercial vehicle standard gearbox with spur toothed sliding gears Rieseler Automatic passenger car trans~ission with torque converter and planetary gear-set Cotal 3-speed planetary gear with automatic shifting via 3 electromagnetic clutches Development of the TRILOK converter - a precondition for modern hydro-mechanical "conventional" automatic transmissions Maybach Overdrive auxiliary gearbox for reducing engine speed; shifting by means of override face clutches, and ground helical cut gearwheels to reduce noise ZF Aphon transmission Helical cut 4-speed transmission with multiplate synchromesh DKW F1 with driven front wheels. Transverse-mounted 2-cylinder 2-stroke engine Wilson transmission Multistage planetary coupling gear with identical ring gears that are alternately fixed against the housing by means of brake bands ZF All-synchromesh gearbox 4-speed gearbox, helical cut, all speeds synchronised General Motors Hydra-Matic transmission First mass-produced conventional automatic transmission: 13 million produced; hydrodynamic clutch, 4-speed planetary transmission, 2 belt brakes, 2 multiplate clutches ZF 4-speed transmission, helical cut, gearshift mechanism with electro-magnetic multiplate clutches GCI1~r(!l !1"fn,'"s T)Y!7(!fJm~'-tr?':!s·

mission with polyphase converter

I

I

and 2-speed Ravigneaux planetary gear-set 1:;50

1950

1952

1953

1953

1958

1961

1961

1962

1962

1962

:

1967

VW Semi-automatic transmission with torque converter clutch and J-sjJt;\"'J g.... al-cJ

PucKuru UitrUfilUlic tfu/tslltissiu/t

lCai-llhjul1lCU

Conventional automatic transmission with torque converter lockup clutch, 2-stage 2-phase converter and 2-speed planetary gear Van Doorne "Variomatic" Mass production of continuously variable V-belt transmission with axially adjustable taper discs (diameter adjustment) Borg-Warner "Warner Gear" transmission Conventional automatic transmission with TRILOK converter and 3-speed planetary gear-set Borgward Automatic transmission with converter and 3-speed spur gear drive with electro-hydraulic shifting ZF Hydromedia transmission for buses; 3-speed transmission with converter and hydraulically activated multiplate clutches Smith Magnetic-particle double clutch with rear-mounted 3-speed spur gear stage transmission and electrically activated dog clutches ZF 3-speed automatic transmission for passenger cars; converter without lock-up clutch, 3-stage planetary gear-set and hydraulic control Daimler-Benz 4-speed automatic transmission, of 2-range design with hydrodynamic clutch ZF 6-speed transmission series for commercial vehicles; dog clutch engagement or synchronised; optional l2-speed version with front-mounted splitter unit (2-unit design) Eaton 9-speed commercial vehicle transmission with power split to 2 countershafts for a shorter overall design length Commercial vehicle range change type transmission designs with 9 and more gears, especially with a rear-mounted range unit of planetary design start to become established

transmission Various companies develop a torque converter clutch transmission for commercial vehicles with a torque converter lockup clutch and secondary 6-8 speed geared transmission Sundstrand "Responder" Mass produced hydrostatic commercial vehicle gearbox with power split through planetary gear-set Turner Commercial vehicle transmission with output constant gear and synchromesh on the countershaft to increase service life Van Doorne Continuously variable transmission for passenger cars with steel thrust chain and axially adjustable taper discs ZF l6-speed commercial vehicle transmission with integral frontmounted splitter and rear-mounted range unit 5-spced passenger car gearboxes with increased overall gear ratio to reduce fuel consumption become established Converter with lock-up clutch in automatic passenger car transmissions Eaton/Fuller Twin Splitter 12speed commercial vehicle transmission with 4-speed main gearbox and 2 rear-mounted splitter units Porsche Re-discovery of the twin clutch principle as an automatic transmission for passenger cars Mass production of conventional automatic transmissions with torque converter, lock-up clutch, five speeds and electro-hydraulic shift Voith Continuously variable hydrostatic power split transmission for buses. Possibility of braking energy recuperation with energy accumulator Renewed interest in alternative power-train concepts: electrical and hybrid drives

I

1970

1971

1972

1975

1976

1978

1980

1983

1985 1990

1990

1991

i

. :,;

R

~ "

~

it ~

1.2 . History of Vehicle Transmissions

1784 Watt patent 2-speed gearbox with dog clutch engagement

13

1821 Griffith 2-speed gearbox with sliding gears

1827 Pecqueur Differential gear

1834 Bodmer Shiftable planetary gear

1849 Anderson Shiftable belt transmission

1879 Selden Complete vehicle transmission with clutch, R gear and housing

1886 Benz Belt-driven bevel gear .differential

Around 1885 Marcus Engaging cone clutch

I-lI------,...----_"!"i-.,-j- - - -

Figure 1.12. Early vehicle gear components and mechanisms

,

---------

With the development of the steam engine, the need atoseto adapt the! enginepoWet available to the intended use. The first steam-powered vehicles were driven by ratchet r;P8f" (Fi~l1re 1,0) Hlghf'r r2.tihs Wr'[P required tr; -:E::18 gra~Ec:nts t1::2cD to Cri\'8 c;-; the flat. In 1784 JAMES WATT patented the constant-mesh gear with constantly meshing gearwheels (Figure 1.12), which is still in common use today. The variable-speed transmission was born. Production of road vehicles only really started several decades later. The steam vehicle builders EVANS and TREVITHICK, 1801, solved the problem of torque adaptation, but still by interchanging a gear pair. .There were a number of important inventions as early as the beginning of the 19th century (Figure 1.12).. In 1821, GRIFFITH disclosed the sliding gear transmission system, which was extensively used as an inexpensive solution into the 20th century~ In 1827, PECQUEUR succeeded in equalising wheel speeds when cornering, by means of a differential. In 1834, BODMER designed a partial power~shift planetary transmission. The change in gear ratio is achieved by disengaging the shifting dogs and tightening a brake band. In 1879, SELDEN patented a sliding gear drive with clutch and reverse gear as part of an overall patent for a piston engine vehicle. It is striking that around the turn of the century there was already intensive effort devoted to the continuously variable transmission, which is ideally suited to the internal combustion engine. This involved considering not only mechanical solutions, but also hydrostatic and even pneumatic solutions (Table 1.4). But they did not gain acceptance, because of their low power rating or mechanical complexity. The hydrodynamic Fottinger torque converter (Table 1.6), invented in 1905 for ship propulsion systems, was not applied to vehicle power trains until 1925.

1889 Maybach-Daimler gearbox

1890 Peugeotdriveline

Figure 1.13. Early vehicle transmissions Direct drive was another important development, with which BENZ created the classic countershaft transmission with coaxial input and output, which remains valid to this day. It is not yet incorporated in the exemplary 1890 Peugeot power train (Figure 1.13). This design of countershaft transmission with direct drive and four forward speeds proved effective in practice. The basic problems of ratio changing were solved. Another phase of development started around 1920. In an effort to improve comfort and ease of use, development effort focused on ground and/or helical-cut spur gears, or reducing engine speed to reduce noise and make changing gear easier. Another important breakthrough was the standard gearbox (i.e. gearboxes which are structurally identical or which vary only in their ratios and connections) to facilitate efficient, cost-effective production (Table 1.4).

1.2

History o/Vehicle Transmissions

15

The first gearshiftingaids date from the year 1915. The ZF Soden transmission had constant-mesh gearwheels, preselector and synchronising mechanisms. This transmission provided preselection, whereby the driver set a knob on the steering wheel to the required gear and pressed on the pedal. The clutch disengages. When the shift pedal is released the pre-selected gear engages. automatically. The advantage of almost effortless shiftin~ could not make up for the dIsadvantages, such as the difficulty of adjusting the cable controls, and the complex gearbox design. In a General Motors transmission, the shifting action and power transmission was effected by means of dogs with a taper synchroniser. In 1928, MAYBACH succeeded in substantially reducing vehicle noise with his auxiliary overdrive and helical-ground gears, by reducing gear hobbing faults and substantially improving the engine speed. At the same time the quiet-running ZF Aphon gearbox was produced, with three gears synchronised with plates. In the ZF fully synchromesh gearbox (1934), all the forward gears already had taper synchronisers. The last striking changes in format without any change in design of mechanical passenger car transmissions occurred after the Second World War, when rear-wheel drive and then front-wheel drive with transverse engine became more prevalent on the market, a development which has now penetrated as far as upper mid-size vehicles. The direct gear and coaxial design were abandoned, and the engine, transmission and differential were combined in one unit to save space. From around 1978, 5-speed geared transmissions with an increased range of ratios and finer ratio stepping became increasingly common. Gear shifting aids, leading up to automatic systems, represent an independent line of development. From around 1956, Fichtel & Sachs supplied DKW (now Audi) with an electrically controlled semi-automatic clutch, the SAXOMAT. The system consisted of a centrifugal master clutch and a vacuum-operated gearshifting clutch. When the gearshift lever is touched, a vacuum-controlled servo device opens the gearshifting clutch. When the gearshift lever is released, air is slowly released to the servomechanism through a nozzle, thus engaging it. Pressing the accelerator pedal increases the flow of air, and thus the engaging action. This represented a considerable advance in ease of operation compared to vehicles with a foot-operated clutch. In 1967 VW presented a semi-automatic 3-speed torque converter clutch transmission for passenger cars. H. RIESELER designed an automatic transmission as early as 1925, consisting of a torque converter and rear-mounted planetary geared transmission, whose main components (torque converter with planetary gear shifted by means of clutches and brakes) are now typical for all automatic transmissions. Rieseler thus made an outstanding contribution, the advantages of which were not yet recognised by subsequent designers. They continued to merely replace the mechanical clutches with a fluid clutch. Conventional automatic transmissions began to establish themselves from 1939, comprising a torque converter (some with a clutch), three- or four-stage planetary gear-set and hydraulic control. The first mass-produced transmission of this type was the General Motors Hydramatic. These transmissions spread rapidly in the USA after the Second World War. They achieved a market share of around 85%. In Europe, conventional automatic transmissions still only reach a market share of around 15% even today. In 1953, Borgward developed the first automatic transmission design in Germany. It had a powershift countershaft transmission with a front-mounted torque converter used only for moving off. After building under licence for some time, Daimler-Benz and ZF launched their own designs in 1961. Daimler-Benz still had an old design similar to the Hydramatic transmission, with planetary gear transmission and front-mounted fluid clutch. These automatic transmissions underwent constant development to reduce fuel consumption. A torque converter lockup clutch and a fourth and fifth gear to increase the

J~_~

--,-

range and adaptation of ratios became standard. ;The introduction of! eleotFohydrau!ic", controls made shifting easier. The continuously vari Vehicle ¢:::> Road Traffic Cargo. This formulation contains a conflict of interest (Figure 2.2). Increasing one's own quality of life is beneficial for the quality of life of society at large only in the short term. If each individual seeks to improve her quality oflife without regard for others, the quality of life of the society in which she lives will suffer. This conflict of interest is starkly illustrated by the current problem of traffic and environment.

.... Cl:l

Q)

>-

E

~

500 -,-------r------,.----,.--,---.,...----:::::::::;;;II.., Specific volume of goods traffic in 10 tonne-km / Year / Person - + + - - - - - 1 - - - - - - 1 400

I

Q) (J)

C C

-

c .Q

0

m

..--

300 -t-------1f-----'--~'__+_-------j-------,._---..,..._i E c

c

c o

()

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....Cl:l ( J)

I i iil '$ ~

i

200 -r---~"'9----t:..-rtI'-- Absolute traffic volume 9

in 10 tonne-km / Year

'U 0 0 0> 0

Q)

:g

I

-- --

100 ~-----:::l#C-+-----I---- Population

-

E :::l

0

> 1960

1970

co

a..

- + - - - - - f - 100

~

-

O-+------+------'--+-------t-----t------t-

1950

:::l

1980

50

o

1990 Year 2000

if; t

FIgure 2.1. bxample: Growth 01 goous tralfic anu popuiclLiou iu GCUllclny; [lgU1~S [v1 the whole of Germany from 1990 [2.1] .

t

Optimum for individuAl and society

Figure 2.2. Relationship between the quality of life of a society and that of an individual Quality of life of the individual

On the subject of the Road Traffic Transport System, H. J. FORSTER writes as follows [2.2]: "Since MAN, with all his wishes and needs, far outweighs all ther interests, optimising the system is not necessarily the same thing as optimising transportati n performance. PeopLe using the traffic system also suffer from its ill effects, both as motori ts and as citizens. CLassic measures of transport effectiveness, such as transportation voLume assenger kilometres), and the cost and speed of traveL,· shouLd therefore become secondary consi rations. Priority has to be given to more compLex human criteria such as journey quality, hu n satisfaction, and especially environmentaL impact. For goods traffic, economic factors such ransportation volume (tonne-km), transport costs (cost per tonne-km) and journey speed (kmlh) c':lorntJ'nue---tfY outweigh considerations of social and environmental impact. "

2.1.1 The Significance of Motor Vehicles in our Mobile World Mobility is an ancient basic human need. There are two factors influencing human choice of means of transport. One is actual satisfaction of objective needs, such as transportation performance, door-to-door access, and attainability of the destination. The other is satisfaction of putative subjective needs such as comfort, convenience, and freedom to decide the mode, destination and timing of the journey. Individual mobility by motor vehicle is also an expression of the freedom enshrined in our social order. Individual traffic is stochastic; it is neither determinable nor susceptible to a planned economy. Public transport is determinable. Its use can be planned. HELLING [2.3] proposes sketching situations and development goals for road traffic by considering it as a black box (Figure 2.3) and comparing inputs and outputs. This simplified view represents the task as achieving the desired transportation output with the least negative side-effects and input of resources. The resources required to manufacture motor vehicles are shown as ambivalent to the extent that they contribute to adding value and creating jobs (Figure 2.4).

1

25

Fundamental Principles of Traffic and Vehicle Engineering Resources consumed

-

Production, maintenance and of vehicles

00 :J

0.~

c

Q)

(tj

> :0 E

Results achieved :E w

....

Production and maintenance of traffic organisation

.

Production and maintenance of roadways (highways and parking space)

«

en > en m ~

.

en 0

Consequences of accidents (deaths and injuries)

u. u. DISTRfBUTfON TRAFFIC

INTERNAL COMPANY TRANSPORT >:

Figure 2.13. Traffic distances: Diagram of a goods transport chain

Because of their lower average speed, scheduled service buses are significantly slower than high-speed urban trains, given the same idle time. For longer trips of around 17 km and more, high-speed urban railways offer shorter travel times than passenger car or taxi. There are five different means of transport for goods: .

o railway, o commercial vehicle (road transport), o ship (canal, sea freight), . o aircraft (air freight), o pipeline. These means of transport often form a transport chain (Figure 2.13). New approaches are urgently needed to reduce the amount of goods traffic on the roads. These means of transport can be compared by various features such as speed of transport, transport flow, space requirement, and transport flow related to space requirement. Transport Transport Profile Means flow surface speed of Cross-sectional transprofile v (km/h) TF (m3/h) A (m 2 ) port

Specific transport flow· TFspec = TF / A (m/h)

Railway

50

20000

37

541

Motorway

50

14500

115

126

Canal

12

6250

470

13.3

Pipeline

7.2

2850

0.4

7125

Figure 2.14. Comparison of goods transport alternatives: rail, motorway, c nal, pipeline [2.4]

i

2.1

33

Fundamental Principles of Traffic and Vehicle Engineering

If these various means of transport are compared in terms of transport speed and transport flow (Figure 2.14) the railway emerges as particularly effective, followed by the truck. The pipeline does not fare well in this comparison. The relative transport flow offers an interesting comparison, showing how well utilised the transport facility is relative to the cross sectional area it requires. On this measure the pipeline is well in the lead, followed by the railway and motorway. Canal navigation emerges unfavourably from this com . son. Figure 2.15 shows the efficiency of various means of transport. The relation hip between overall weight and payload is best with the pipeline, followed by the b ge, railway and truck. The payload ratio is much less favourable for a~rcraft.

8-,----------------------, "0

ro 6 0 >. ro c.

:;:,

..c: .2' 4 Q)

3: ro

•• E3

Pipeline

6.7

Barge Railway

[S) Truck

an

Aircraft

~

Q)

> 0 2

1.3

1.4

1.7

o Figure 2.15. Relationship between total weight and payload for various means of transport [2.4]

Commercial vehicles carry most of the annual volume of goods traffic (Figure 2.16), with rail and barge well behind. There is no immediate prospect of sufficiently expanding rail freight to noticeably relieve the burden of goods traffic on our roads. A key feature for a goods transport system is door-to-:-door access -:" enabling goods to be transported by the same means of transport without transhipment. The reason for the enormous increase in the number of trucks is door-to-door transport, speed and economic efficiency, and just-in-time delivery to assembly lines. Transport systems have to be assessed on the basis of satisfying transportation needs, environmental impact, and energy efficiency. First of all certain structural adjustments have to be made. In particular road and rail must be treated equally in financial terms. In this respect rail is at a disadvantage to road traffic. Using the road as a cheap storage facility in the just-in-time system of delivery is not economically viable in the long term. It contributes to traffic congestion.

f I

2.1.4 Alternative Transport Concepts Innovative mass transit systems have been under consideration since around 1960. There is a distinction between local transport within conurbations, and high-speed links for easinp- the hurden of long-distance traffic on the roads. Prototypes of such concepts are in 1.·.. •.

._....,-e_x_is_t_en_c_e, some of them using new technologies such as maglev or an cushion lechnoiogy. Some experimental tracks have been constructed.

~

~

I

480

...

440

(Ij Q)

>- 400 -. E

~ I

Q)

c 360 c

0

0>

a

320

~

.~o

c (Ij

280 I-r--I-i-~~~~=~"'-Ir---:-:r'l--n~

. E 240

...

.g ~

200

-t--+---t---±--~"f--~lIIiIlIIl~-+---+---#+---=,~

-+---t--~T_--I_-__+--_+--_+_--+_lr__i ......

c

o

~ t: o a. ~

-

15%

160 +--~-_+_-~~=_=_::_=:__:_._:_~~::I_~~~~__l: 120

Railway

-+---t---+---I_---:±_---t--'---_+_---+--~

0.2% ) ( Aircraft

~

~

80

s

40l::t::.~~--~~~~~~;::::::E::::t==t=J o

-+---t-~0fIIIIE'%::-::=___-1_-__+-_'_i~ _ _..L

1966

1970

1974

1978

1982

1986

1990

Year

997

Figure 2.16. Growth of goods traffic in Germany

These concepts have not s far achieved a brea rough, offering only minor advantages over conventional transp rt systems. This a ies all the more since the railway also has further development po entia!. The only one of t ese systems whi appears to be establishing itself is the TRANSRAPID system [2.6] This is a magI v monorail with linear motor drive. Designing and developing such syst ms is expe sive, and has to be co-ordinated internationally to make it viable. Their success s on legislation and market acceptance. Both conventional and innovative vehicles and transport systems for local transport can be categorised in a morphological table by control system and type of use (Table 2.2). Methodical analysis of systems along these lines can help to develop proposals for innovative transport systems [2~7 -2.14]. . Buses are considered to have better than average development prospects for local and regional transport because of their flexibility in use, the low level of investment required and relatively low energy requirements. The significance of buses has been reflected in intensive development of automatic transmissions for buses. An interesting new development is buses that are partly guided along tracks, operating in "dual mode" (Figure 2.17). This means buses can operate both freely on conventional roads and also under guidance along special tracks. The benefits of such systems are reduced driver stress,tracks that are easy to build and less environmentally damaging, and minimal tunnel diameters. "Demand buses" are mini-buses which can be called to bus stops. A process computer optimises transport routes within the service network and notifies the passenger of his time of departure.

"

2.2

The Market and Development Situationfor Vehicles, Gearboxes and Components

35

Table 2.2. Classification of transport systems by control system and type of use. (Individual utility decreases from top to bottom) [2.3] Control

Use Individual

Free

Passenger car Commercial vehicle Motorcycle

On demand

Planned

Taxi Demand bus

Schedule service bus

Dual mode

Automated motorway

Dual-mode taxi

Dual-mode bus

Track-bound

Transport belt

Cabin taxi

Railway

Having lost the initiative to road traffic for decades, railways such as Germany's ICE, France's TGV, and Japan's SHINKANSEN are using high-speed trains which substantially reduce rail journey times. A European network of high-speed trains is emerging, with speeds of 250 km/h and more. These developments are characterised by high transportation performance, virtual door-to-door access, and a high level of passenger comfort.

Vehicle with electronic guide track system Stationary guide cable transmitter Transmitter 1 Transmitter 2

Figure 2.17. Bus Front reception aerial with automatic electronic guide-rail conin vehicle . trol. Source: MAN Three-core guide cable

2.2 The Market and Development Situation for Vehicles, Gearboxes and Components

t

i

A progressive vehicle and gearbox development process must be not only technically . sophisticated but also market-orientated. Vehicles and vehicle transmissions are developed cyclically, and have a relatively extended product and production life-cycle. Vehicle transmissions generally only require redevelopment after some 10-15 years. The transmission developer must therefore be familiar with the market situation, and be able to assess the market and changing values in society in the long term. This requires continuous observation of the market and of technological developments, as well as project !1hmninr: i1T1~lpmpnt~tinn nnn (m~l~,(Ci" nf "flltnri(Cti~" rrnjp,rt". Tllr n lTe r t ~r(\(hlrt npv,=,lnr.-.. - ment .sions generally lead to serious financial loss.

~

------r--__._.....;;..,_--,-

-.i--";""--O

,

2.2.1 Market Situation and Production Figures ; ~:l\.; YL-hidc .i.nLiusll)' IS dU I111fJVlldHl facLv.l 111 LIle

glvbal \.;COHVH y. ,)3.0 w.i..ilion illVLOl ve-

hicles were produced world-wide in 1997 (Figures 2.18 and 2.19). This figure comprises 38.6 million passenger cars and 15.2 million commercial vehicle [2.5].

Definitions: Passenger car: Motor vehicle designed and equipped ma'nly for transporting . people, with a maximum of nine seats. ~ . ~ / Commercial vehicle: Motor vehicle designed for the purpose of: transpo~ - Bus; for transporting goods and pulling trailers - Truck; or just for pulling trailers - Tractor. This excludes passenger cars. There are three competing centres of motor vehicle development: Europe, the USA and Japan/South Korea. . Europe is the largest producer of passenger cars (Figure 2.18). The proportion of passenger car production accounted for by sub-compacts (up to approximately 1500 cc), mid-size passenger cars (up to approximately 2500 cc) and luxury passenger cars varies greatly in the various European producer countries. Whereas France and Italy produce mostly sub-compacts and mid-range passenger cars, Germany produces a larger proportion of mid-range and luxury passenger cars. Germany produces more luxury passenger cars than the rest of Europe put together. The USA and Canada are the largest producers of trucks over 2 t (Figure 2.19). Each market has specific conditions dictated largely by the economic and social circumstances of customers, social values, geographical factors and, importantly, legislation. Motor vehicles must satisfy market requirements to be successful. This affects the gearbox in particular, as the link between the engine and the road. Whereas in the USA more than 80% of passenger cars are fitted with automatic transmissions, the figure in Europe is only 15% (Germany 20%). In economically prosperous countries there is a market for ever better equipped passenger cars, for example with power assisted steering, automatic gearbox, air conditioning, airbag, antilock braking system, etc. Gearboxes for commercial vehicles over 4 t gross weight are selected specially for the particular application. There are often different numbers of speeds and different methods of operation (manual, semi-automatic or fully automatic) available for a commercial vehicle gearbox from different manufacturers. The spectrum of types of transmission for commercial vehicles is as broad as the spectrum of applications. In the USA for example, constant-mesh gearboxes are mostly used for trucks weighing over 16 1. For long stretches where no shifting is necessary, the driver is equipped with the less convenient unsynchronised constant-mesh gearbox. This also applies in many developing countries, where driver comfort is of less concern than the longer service life of the constant-mesh transmission. In Europe on the other hand synchromesh gearboxes dominate for heavy trucks as well, accounting for about 60%. The trend is towards semi-automatic or fully automatic synchromesh gearboxes. Much European truck production is in the class up to 4 t gross weight rating (Figure 2.19). These vehicles normally have 4-speed or 5-speed synchromesh gearboxes. These are often similar to passenger car gearboxes, or are modified passenger car gearboxes. Assuming that 10% more gearboxes than vehicles are produced, to allow for spare parts, it is possible to estimate the number of transmission components comprising gears and synchroniser packs. In Europe in 1997 approximately 160 million gearwheels and 37 million synchroniser packs were manufactured for passenger car synchromesh gearboxes. Approximately 30 million gearwheels and 6 million synchroniser packs were produced for commercial vehicle constant-mesh transmissions and synchromesh gearboxes.

r

2.2

The Market and Development Situationfor Vehicles, Gearboxes and Components en 60

c

.Q

·E

50

oS 40 c o :;:::; o 30

e ::J

a.

20

~ 10 o

~

0 -f!==::r==;::===F=:;= 1976

b] Luxury from 2.5 I •

Mid-range up to 2.5 I

mSmall car up to 1.5 I

Germany

Europe

Others Canada

o

20

4 speed

Europe 1997: Car gearwheels

t_---,.

40

60

100 %

80

5 speed

Conventional auto-

Chain converter

matic gearboxes

gearboxes

Europe 1997: Car synchroniser

produced

FIgure 2.18. Production figures or passenger cars

packs produced 37 mill.

37

\

\ (/) 60 c

.2 SO

·E

-

.£ 40

( /)

c:

go

cu

o ~ o

30

::J

-c

K20

0

u..

1997 Total

~ 10 ~ 0 ~::::.::::;i:;i::::i~c: : o: : m:;: m: : e: : rc: : ia: ;:l: : v: : e: : hi: ;cf: : e: : s: : :i 1976 (/)

c .Q

·E

18

D

.£ 16 c

Up to 4 t

go 14

~4t-8t

ec- 10

(SJ8t-16t

-5

12-

O>

8

D D

From 16 t Tractor units

~ 6 ~ 4 .~ 2

~ 0 ~:::::J!Z:=;;:Z==;:~lilllilllJ:::l:~ill.llillll!::::d::::WJJlJ.LlJ!~::::l.llUill.lJlb'

E o ()

Germany

Europe

USA & Canada

Others

100 %

Automatic gearboxes

Europe 1997:

Europe 1997:

Commercial

Commercial vehicle

vehicle gear-

synchroniser packs

wheels produced

produced

Figure 2.19. Production figures for commercial vehicles

Others

6 mill.

2.2

The Market and Development Situation for Vehicles, Gearboxes and Components

39

These figures illustrate the enormous economic importance of the motor vehicle. No other product supports the production of such technically sophisticated components in such quantities. There is no product in sight, which could replace the motor vehicle as the engine of the economy. Table 2.3 shows Germany's balance of trade in motor vehicles in recent years. 3.04 million motor vehicles were exported and 2.15 million imported in 1997. Table 2.3. Germany's balance of trade in motor vehicles [2.5]

E

X P

0 R T I

M P

0 R T

Quantity in millions

1991

1992

1993

1994

1995

1996

1997

Passenger cars

2.20

2.57

2.06

2.27

2.47

2.65

2.82

Commercial vehicles

0.16

0.16

0.12

0.14

0.17

0.19

0.22

Total

2.36

2.73

2.18

2.41

2.64

2.84

3.04

Passenger cars

2.52

2.18

1.64

1.64

1.77

1.85

1.93

Commercial vehicles

0.22

0.21

0.11

0.15

0.16

0.19

0.22

Total

2.74

2.39

1.75

1.79

1.93

2.04

2.15

Table 2.4 lists the main independent manufacturers of motor vehicle transmissions. Vehicle transmissions, especially mass-produced passenger car gearboxes, are mostly produced by the motor vehicle manufacturers themselves. Table 2.4. Some independent manufacturers of vehicle transmissions Passenger car mechanical

Japan

Commercial vehicle mechanical

- GETRAG -ZF

-ZF ~ -ZF - BORG- EATON WARNER - GENERAL MOTORS

-AISIN - NEW VENTURE GEAR

-AISIN - BORGWARNER

-AISIN

-AISIN -JATCO

Western Europe

USA

Passenger car automatic

1\

Commercial vehicle automatic -ZF - VOITH -RENK

-:~A

- ALLISON (GM) -TWIN DISC

-AISI'-J

-AISIN -JATCO

-EA ON - NE I VENTURE GEA R -CLA RK

2.2.2 Development Situation The pace of technological development has accelera~ d in recent years. Microelectronics ~'c;1tilJ.LIC:; to fi;~J ;iC\V ~hj:il..-atiojiS lilv"..J,iclc::i, vcliiSlv Lftiii~lHis~ilJ!L'; ctllJ iii L~lcil Jcvciup-

I

I

-l.--.,...-----,..-------"--~---.-_r__-----·

ment. The pace of product development is increasingly becoming an impbrHm:t competitive factor for individual companies. Modem GCVd0tlHl-';liljJlvC-';.)0C,::, slI'll

w

Figure 2.20. Definition of the subject area of vehicle transmissions

Electrics/ electronics

2.3

Basic Elements of Vehicle and Transmission Engineering

41

2.3 Basic Elements of Vehicle and Transmission Engineering It is essen~ial to clearly define the vehicle and its intended use as the starting point for targeted vehIcle gearbox development. Definitions and basic physical elements of automotive and transmission technology are explained below. They are the basis for the observations in the following chapter. The topic of "Automotive Transmissions" in this book covers all components of the power-train assembly with the exception of the engine (Figure 2.20). In the development of vehicle transmissions a distinction must be made between variables which the designer can influence (internal factors) and those he cannot influence (external factors). These factors are set out in Table 2.6. Table 2.6. Internal and external factors affecting the development of vehicle transmissions Internal factors which can be influenced by the design engineer

o

Bodywork

External factors which cannot be infl~enced by the design engIneer

0 Road profile

0 Chassis

0 Driving style

0 Electrics/electronics

0 Payload

0 Engine

0 Traffic conditions

0 Vehicle transmission (see Figure 2.20)

0 Weather conditions

2.3.1 Systematic Classification of Vehicles and Vehicle Use The development of a vehicle transmission relates to the type of vehicle, its power unit and its intended use. A classification of vehicles oriented to transmission development assists systematic analysis. Table 2.7 shows a transmission-oriented classifica~'on of vehicles which has proved effective in practice. Vehicles are first divided into pa senger cars, commercial vehicles, construction vehicles, tractors and special vehicles. Pas 'enger cars are split into two main groups by engine size: up to 75 kW and over 75 kW. \ The commercial vehicle category is split into buses land trucks. The truck category is further broken down by gross weight. \ The bus category can conveniently be broken down by function into urban and local buses, and long-distance coaches. \ . A further relevant dimension is the type of use intended. here are three baSIC types of use in a transmission-orientated classification of automobiles:

o o o

On-road use. On/ojj~road use, e.g. construction vehicles. This combined type of use, w i typical e.g. for dump trucks, means the transmission must provide economical propulsion both on and off-road. Off-road use. Vehicles move predominantly off-road, possibly with occasion~l on-road use. This category includes tracked vehicles or extremely heavy specIal vehIcles not pernutted on normal roads, such as lanofill velncles or minll1g vehicles.

I -L--.,----r-----------r--------,.-----

Table 2.7. Transmission-oriented classification of motor vehicles by type of vehicle and type of use; GVW: Gross vehicle weight. *) Another feature is the number of seats Type of use

Type of vehicle

Power P < 75 kW Power P > 75 kW C/)

Q)

~ ~

I

/

.-.

1-_~=L::::iQ=h=t=co:::::m=-m~e_rc_i-.:...a_1v.:. . e.:. . h.:. . ic.:. . l:. .-es.:. . :_G.:. . V-.:. .-.:. .W:. -.

()

c: 96 Q)

'0

~

Figure 2.33. Efficiency of a mechanical commercial vehicle gearbox for various speeds where nl = 1500 lImin and T) = 500-2000 Nm; calculation as described in [2.21]

w

94

1st gear 4th gear 8th gear 8th gear slow slow slow fast (direct)

2.5 Transmission Design Trends The range of transmission designs has become much more diverse since around 1975, especially in the case of passenger car transmissions. 4-speed synchromesh gearboxes were standard, as were conventional 3-speed automatic transmissions with a small market share of approximately 15% in Europe. The Figures 2.34 and 2.35 show the variety of emergent designs. The 5-speed synchromesh gearbox is now standard. 4-speed and more recently 5-speedautomatic transmissions will increase their market shares somewhat. But they will encounter increasing competition from continuously variable transmissions, which are to be available from 1995 in mid-size vehicles. But this is conditional on continuously variable transmissions really delivering the fuel economy they promise.

%

O..J.JJ:..LliJw.LWJ..u.u..u.u..L.l.f-L.L.L.Ll..LU..J..LU.Jw.LUllWjlWJ..L.I:

1980

1985

1990

1995

Year

2000

Figure 2.34. Trend in the use of passenger car transmissions in Europe [2.8]

2.5

Transmission Design Trends

57

Today

4/S-speed electrohydr. automatic transmission

CVT Continuously variable transmission 6-speed electrohydr. automatic transmission

Figure 2.35. Passenger car transmissions today and in future The problem with automatic and continuously variable transmISSIOn remains that of broad acceptance by customers. Most drivers still prefer personal control, being actively involved in the process. 6-speed synchromesh gearboxes will only occupy a niche market for sports cars and high powered Diesel cars. As average speeds on highways fall, such manual transmissions are not very appropriate, especially since the frequency of gear changing will increase. Electronic clutch controls for dry clutches are currently about to enter mass production. They can be combined with conventional selector gearboxes. Automated multispeed gearboxes (8-speed transmissions of double range construction) may provide an alternative to continuously variable transmissions for passenger cars. Mechanical gearboxes with five to sixteen gears of single- or multi-range design will remain the standard for commercial vehicles, and will increasingly be available in semiautomatic or fully automatic form to reduce driver stress. The shifting process is already semi-automatic. Automating moving off with automated dry clutches instead of torque converters is difficult to achieve with commercial vehicles, but the problem has in principle been solved. Conventional automatic transmissions with torque converter, lockup clutch and planetary gear-sets have only established themselves in buses, which will remain their domain. The success of hydrostatic automatic transmissions for commercial vehicles and electrical drives for buses remains to be seen.

J__---,--

,

-

3 Mediating the power Flow

Reciprocity: Supply and Demand

Vehicle transmissions mediate between the engine and the drive wheels. The transmission adapts the power output to the power requirement by converting torque and rotational speed. The power requirement at the drive wheels is determined by the driving resistance [3.1, 3.2].

3.1 power Requirement The anticipated driving resistance is an important variable when designing vehicle tranSmissions. Driving resistance is made up of

o wheel resistance FR' o air resistance FL' o gradient resistance F St , and o acceleration resistance Fa . 3.1.1 Wheel Resistance Wheel resistance comprises the resisting forces acting on the rolling wheel. It is made up of rolling resistance, road surface resistance and slip resistance.

Figure 3.1 shows the forces and torques acting on the wheel. The integral of the pressure distribution over the tyre contact area gives the reaction force R. It is the same as the wheel load GR. Because of the asymmetrical pressure distribution in the wheel contact area of the rolling wheel. the point of application of the reaction force R is located in front of the wheel axis by the amount of eccentricity e.

b)

a)

Figure 3.\. Furces and torques at rhe wheel. a) stretches

00

the level; b) on uphill/downhill

3.1

power Requirement

59

.If the wheel is unaccelerated and driven by TR, then TR = Furdyn + R e .

(3.1)

for a wheel rolling without drive torque and braking torque (TR =0) e -Fu = - R .

(3.2)

rdyn

.. The circumferential force - Fu is equal to the rolling resistance force FR R II given these assumptions. On a level surface R = GR, and therefore ' 0 (3.3)

Trials have revealed an almost linear correlation between the rolling resistance force FR. Roll and the wheel load GR· The relationship is defined by the formula (3.4) The dimensionless proportionality factor fR is designated the rolling resistance coefficient. From (3.3) and (3.4) it is given as (3.5)

Table 3.1 shows standard values for rolling resistance coefficients both on and off-road. Rolling resistance is chiefly a function of ground speed, wheel load, tyre pressure and tyre type. The influence of speed can be described according to [3.1] by the following polynomial: .fR = Ao + A I V + A2 v 4 . Since driving resistance calculations normally assume straight running on a dry surface, and rolling resistance is anyway the dominant wheel resistance, wheel resistance is normally assumed to be equal to rolling resistance. The following formula then applies: (3.6)

Table 3.1. Reference values for the rolling resistance coefficientfR. For road speeds below 60 km/h,fR can be assumed to be constant. (See also Table 5.1) Road surface Firm road surface Smooth tarmac road Smooth concrete road Rough, good concrete surface Good stone paving Bad, worn road surface Unmade road surface Very good earth tracks Bad earth tracks Loose sand

Rolling resistance coefficient fR

0.010 0.011 0.014 0.020 0.035 0.045 0.160 0.150-0.300

o

lise;"

d

When travelling up gradients/down slopes at an angle of aSt (Figure 3.1 b), then . (3./ )

For the w?ole v~hicle wi.th a mass mF, the wheel resistance FR, which is considered equal to the rollmg reSIstance, IS thus given by (3.8)

In the l.ower speed range, the rolling resistance coefficient can be regarded as a constant at. the fIrs~ approximat.ion. The angle of inclination aSt can be ignored on normal journeys WIth gradIents/downhIll slopes of less than 10%. With a gradient of 10% aSt ~ 5.7 0 and thus cos aSt ~ 1.

3.1.2 Adhesion, Dynamic Wheel Radius and Slip There is a frictional connection between the tyres and the road surface. The transmittable circumferential force Fu, (Figure 3.1a), is proportional to the wheel load reaction force R, with a maximum value F u, max = Fz, max =,uH R .

(3.9)

The maximum traction Fz between the tyres and the road surface is constrained by the adhesion limit (Figure 2.22). See also Section 6.1.4 in relation to circumferential force, lateral force and Kamm circle. Table 3.2 gives static friction figures ,uH of pneumatic tyres on road surfaces. Table 3.2. Static coefficient of friction PH of new pneumatic tyres on road surfaces [3.3] Road speed

Static coefficient of friction PH

km/h

Dry road surface

Wet road surface

50

0.85

0.65

90

0.80

0.60

130

0.75

0.55

For many vehicle dynamics calculations, the radius of the tyres is needed (Table 3.3). A distinction is made between

o o

static wheel radius rstat

The distance from the centre of the wheel to the datum plane with the wheel at rest, dynamic wheel radius rdyn

.

.

Calculated from the distance travelled per revolutIOn of the wheel, rollmg without slip. The dynamic wheel radius is calculated from a distance travelled at 60 km/h [3.4]. The increasing tyre slip at higher speeds roughly offsets the increase in rdyn.

3. J

Power Requirement

61

Table 3.3. Dynamic wheel radius of common tyre sizes [3.5] Size

Rolling circumference

rdyn

(m)

(m)

Size

Passenger cars

Rolling circumference

rdyn

(m)

(m)

Passenqer cars

135 R 13

1.670

0.266

205/65 R15

1.975

0.314

145 R 13

1.725

0.275

195/60 R15

1.875

0.298

155 R 13

1.765

0.281

205/60 R15

1.910

0.304

145/70 R13

1.640

0.261

155/70 R13

1.680

0.267

185 R 14

1.985

0.316

165/70 R13

1.730

0.275

215 R 14

2.100

0.334

175/70 R13

1.770

0.282

205 R 14

2.037

0.324

175 R 14

1.935

0.308

195/75 R16

2.152

0.343

185 R 14

1.985

0.316

205/75 R16

2.200

0.350

195/70 R14

1.940

0.309

185/65 R14

1.820

0.290

12 R 22.5

3.302

0.526

185/60 R14

1.765

0.281

315/80 R 22.5

3.295

0.524

195/60 R14

1.800

0.286

295/80 R 22.5

3.215

0.512

195/70 R15

2.000

0.318

215/75 R 17.5

2.376

0.378

185/65 R15

1.895

0.302

275/70 R 22.5

2.950

0.470

195/65 R15

1.935

0.308

305/70 R 19.5

2.805

0.446

Liqht commercial vehicles (vans)

Trucks/buses

The slip between the tyres and the surface can be described as drive slip

UJR rdyn - U F

ST =

,

(3.10)

,

(3.11 )

UJR rdyn

brake slip where

UF

U F - UJR rdyn

Sa

=

uF

is the actual vehicle speed.

3.1.3 Air Resistance

t.

Air flow around the moving vehicle, and through it for purposes of cooling and ventilation. The air resistance is made up of the pressure drag including induced drag (turbulence induced by differences in pressure), surface resistance and internal (through-flow) resistance. The air resistance is a quadratic function of the flow rate. The flow rate U is derived from the sum of the vehicle speed UF and the wind speed component Uw in the direction of the vehicle longitudinal axis. If the wind speed direction is the same as the direction of travel of the vehicle (following wind), then the wind speed is deducted from the vehicle speed to calculate the flow rate. Driving resistance calculations normally assume calm, in which case: U = UF. Air resistance is calculated from the product of dynamic pressure Ih PL U 2 and the maximum vehicle cross-section A multiplied by the dimensionless drag coefficient CWo /\t ~~n ~::- pl·C"2:.:r: Cof l.en3 h",:', :'. I"cbt;'/c 8;:' !-'.lJnl:c';ly of hQ% [In:l (1 t(':ITJl1PT8tprf' ()f In 0(' YP L th_e_ai_r_d_e_n.S.i_t___ _=.,._1._199 k,...g...../m-"'""3.

_

Table 3.4. Reference values for drag coefficient cwo In the case of' goods trucks, the Cw coefficient and the maximum vehicle cross-section are very much dependent on the particll1ct!'

0 P c;irrn '-"

Vehicle

Motorcycle with rider BMW R100S

ew A

ew

A (m 2 )

0.5-0.7

0.7-0.9 0.79 1.7-2.0

0.85-1.4

1.7-2.3

0.4-0.9

2.0

0.60

2.0

0.44

6.0-10.0 7.4 6.0-10.0

2.4-8.0 3.33 2.7-8.0

0.59

(m 2 )

0.4-0.6 0.47

Open convertible

0.5-0.7

Limousine

Truck (solo)

0.22-0.4 0.30 0.22 0.4-0.8 0.45 0.45-0.8

Truck with trailer

0.55-1.0

6.0-10.0

3.3-10.0

Articulated vehicle

0.5-0.9

6.0-10.0

3.0-9.0

VW Golf, '92 model Compact car "2000"

Coach Kassbohrer Setra S315HD

The drag coefficient Cw represents the special case of straight flow, i.e. the wind direction is in line with the longitudinal axis of the vehicle. Table 3.4 gives the cw-values and the maximum vehicle cross-sections (projected frontal area) of some vehicles. Drag is calculated by I

FL =- PL Cw A v

2

(3.12)

2

The aerodynamics of vehicles with high-drag flow-impeding bodywork, such as commercial vehicles, can be greatly improved by the use of air dams.

q = tan lXS t

Vertical projection

= Horizontal projection

q'= q ·100 in % c

o

'';:;

u

Q)

...0-

'0 ~

u

t:Q) > Horizontal projection Figure 3.2. Forces acting on the vehicle travelling uphill

.1

3.1

Power Requirement

63

3.1.4 Gradient resistance The gra~ient resistance or down~ill forc.e relates to the slope descending force (Figure 3.2) and IS calculated from the weIght actmg at the centre of gravity:

(3.13) The road gradient q is defined as the quotient of the vertical and horizontal projections of the roadway (Figure 3.2). When designing roads, gradients of more than 7% are normally avoided. Except in extreme cases, the following approximation is valid

. sm aSt

~

tan aS t

q'

=-

100

.

(3.14)

Table 3.5 shows the maximum gradients (q'max) of some Alpine passes. Table 3.5. Maximum gradients of some passes in Europe

Pass

q'max

Germany: Achen pass Oberjoch France: Col de Braus Iseran Italy: Brenner highway Stilfser-Joch

14% 9% 15% 12% 12% 15%

Pass

q'max

Austria: GroBglockner Timmelsjoch Turracher H6he Wurzen pass

12% 13% 26% 18%

Switzerland: Simplon St. Bernardino St. Gotthard

10% 12% 10%

3.1.5 Acceleration Resistance In addition to the driving resistance occurring in steady state motion (v = const), inertial forces also occur during acceleration and braking. The total mass of the vehicle mF (translatory component) and the inertial mass of those rotating parts of the drive accelerated or braked (rotary component) are the factors influencing the resistance to acceleration: Fa

= mred, i a,

m red, i = m F +

with

(3.15)

LJred,i

(3.16)

2

rdyn

The rotational component is a function of the gear ratio. The moment of inertia of the rotating drive elements of engine, clutch, gearbox, drive shaft, etc., including all the road ",..,1", .• ---1,.-, (,....\,.,-,,-.., .. , l.J."-'CJ.J ,'"-'

I

L-.._ --.-

'y

+l,n

L.L.. I...I..lL

r,"'l~

"",,:,:,'~

.....

~ 2 -l----l----",ollPl.~n o

a..

I:Sl

Synchron.

II

Seal

Ed

Bearing

~

Lubrication

[]

Gearing

o-f--l~..l.-J...J..t_l-W.....L.l.+L...LL.L-W+J.....Ll...LLJ.,..L.JU-l-L.Lll 1000

2000 3000 4000 Gearbox input speed

1/min

6000

nG

~igure In

3.5. ~istribu~ion and size of power losses of a coaxial 5-speed manual gearbox 4th gear (dIrect dnve) at maximum engine torque [3.6] ,

or equivalently the power losses Pv of the individual components of the power train (Figure 3.5):

o

gearing losses: - friction losses, load-dependent, - churning and compressing losses attributable to splash lubrication, unrelated to load,

o

bearing losses: - friction losses, load-dependent, - lubrication losses, unrelated to load,

o

sealing losses: - friction losses caused by radial shaft seals at shaft exits, _ friction losses caused by piston rings used to keep oil under pressure at the shift elements, synchronising losses: _ fluid friction between synchroniser ring and friction taper,

o o

clutch losses: _ fluid friction with wet running, multi-disc clutches and brakes in automatic gearboxes and automated manual gearboxes,

o

torque converter losses: - losses in the torque converter,

o

auxiliary units: - power to drive auxiliary units.

The main losses involved relate to the following components of the power train: o moving-off element, e.g. torque converter, o selector gearbox, e.g. gearwheel transmission, pulley transmission, o final drive, o auxiliary units, e.g. steering pump, oil pump in automatic gearboxes, air conditioning system, variable displacement pump in continuously variable transmissions.

3.1

Power Requirement

67

A further distinction is made between losses that are a dependent on the input speed and the input torque, a dependent only on the engine speed, including in particular pumps directly driven by the engine, a almost independent of rotational speed and torque. For example the efficiency of the final drive is normally assumed to be constant. Figure 3.6 shows the efficiency of the power train from the engine output shaft through to the drive wheels in fourth gear of a 5-speed manual gearbox. Provision is made for a steering pump as an ancillary unit. The level of efficiency only declines rapidly in the low load area. In the case of manual gearboxes, a constant level of efficiency can often be assumed with sufficient accuracy. In the case of continuously variable transmissions, the part-load efficiency is significantly worse, and the drag torque is significantly higher (Figure 3.6).

100 ~--'''''---''--''''''-~-~---r--;--r----r---, Input speed: 3000 1/min

0/0

2

80 -+==-=~="'::I=-;.c.+--+--+---t--+--t--r----j

~

>.

g Q)

60 -W---I---t-----+~

'0

::E Q) Q)

40 -1-1---1---+-----+---1

.~

~ .;::

20 -u--+-~+----+--l

o

5 o-IL~I--.--A-~..j........,J-.,.~...,..._........,____,...:....,~T_r__r_,___j

o

1

. . ,.-

'np:..!t tcm:;ue T

Figure 3.6. 3-D total power-train efficiency map of the direct 4th gear of a 5-speed manual gearbox. The effect of a power steering ~~~~r:~l:

is 1nc!~~dcd

3.2 Diversity of Prime Movers llle ulivil.lg le~.1~LaIlCe Jeselibeu 10 See[ion J. i has LO be overcome oy lhe pnme mover 111 co-operatIon WIth the other components of the power train. " T~,e energy supply, whic~ has ~o be c~rried in the \'ehicle, means "dead" weight and ~ead volume. Energy su~phes wIth as hIgh an energy density as possible are desirable. FIgure 3.7 shows the workmg capacity at the drive wheels from various types of eneroy supply. t:: MJ

MJ I

kg a) Related to mass

Q) Q)

Q) Q)

..c 7

..c 7 -

~

Q)

~ 6, +-'

+-'

+-'

>.

5

ctl

>. +-' .~

+-'

'0 ctl 0. ctl

u

4

0.

L-

0

1

0

~

i I

5-

4-

~ 3en

3

en c :52 2 ~

I I I I I b) Related to volume

~

6 ..c +-' ctl

I

'0

a> en Q)

B

- en

0

'-

C

0>

Q)

a..

""6

CI3

CI3

:Q

.s::

::J

W

0-

:.:i

""6 c C'I:l .s:: .-

C\l

J:

en

::J

Q)

0

~

en CI3 CJ

Q)

C\l

J:

-0

'S 0-

::i

C\l

J:

"0 .;::

..a >, I

:gc

2-

~

1-

~

.....Q)

CI3

..a "0 '(3

o

"

I

.-o

e

en

en

a..

:Q

mal Q)

i:5

n

g

C'I:l O>C'I:l ::J

0-

.s::

W

I

AIcohol

IIn C\l

J:

CI3 -.... "0

r->

>,

2

:g ro .0 ..a >.

I

::J

"0 '(3 C'I:l

=0

CI3

CI3

Hydrocarbon

C\l

Q)

Q)

...J

.....J

I Hydrogen I Electricity

Hydrocarbon

I

AIcohol

I Hydrogen I Electricity

Figure 3.7. The energy available from various forms of automobile energy storing systems. Mechanical energy available at the drive wheels related to: a) mass and b) the volume, of energy supply + energy accumulator (container). Various levels of engine efficiency with energy conversion are taken into account [3.7] The weight of the energy accumulator is factored in, as is the transmission efficiency figure (energy at the wheel/energy of the fuel). Further important criteria in selecting transportable power storage are rapid recharging of the energy accumulator and the necessary infrastructure. Diesel oil and petrol fare best in these respects. The space required to stow batteries is about 30 times greater than that required for diesel fuel or petrol with the same capacity. Fuel cells are however now approaching the capacity of diesel fuel and petrol where H2 is extracted from methanol on the move. Source: VDI-Nachrichten, No.: 22, 30. Mai. 1997.

3.2.1 Overview The drive system of a vehicle can be made up of a variety of combinations of components for storing energy, converting energy and converting output. The prime mover used is a crucial factor determining the assemblies and design of the associated power train.

3.2

Diversity o.lPrime Movers

69 P RIM E MOVERS

Combustion engines

~

Hybrid drives ~

Electric motors

I Internal

r--

External

to-

combustion

combustion

Continuous

Single ignition

combustion Spark ingnition engine f - Diesel engine L - Wankel en gine f-

L Gas turbine

t-

Stirling engine

to-

DC motor

t-

Lear Delta engine

r-

AC motor

'---

Steam engine

--- 3-phase motor

Figure 3.8. Overview of prime movers for motor vehicles; AC: alternating current; DC: direct current

Various prime movers could be used in a vehicle (Figure 3.8). They can be broken down into. combustion engines an? electric motors. In selecting a suitable power unit, the followmg factors must be consIdered: o operating performance drive characteristics, ease of control, startability, energy accumulator, etc., o economy specific energy consumption, specific manufacturing cost, etc., and o environmentfriendliness pOllutant emissions, noise, vibration, etc. In the case of the electric motor it should be noted that there is as yet no satisfactory solution to the problem of energy transport in the vehicle. The fuel cell is currently the most promising technology in prospect. The engine characteristic is a decisive technical consideration in selecting the prime mover, i.e. the power available at full load across the engine speed range. The operation of the various prime movers is not discussed here. References are made to the literature as appropriate. In the discussion that follows, the term internal combustion engine refers to a spark ignition or diesel engine.

3.2.2 Electric Drive The direct current (DC) motor has the ideal motor characteristic where P = const, which corresponds to the ideal traction hyperbola. It can be operated from rest, i.e. motor speed zero. 3-phase motors (synchronous and induction type) offer the advantage of being smaller and lighter than DC motors, making the preferred choice for automotive applications. Figure 3.9 shows the components of an electric drive. It is not essential to have a speed/torque converter. The use of passenger cars with exclusively electric drive will remain restricted to cit]·c.('-

'.J

"l~".-l

L..t.il\....~

tl,n~ ..... (..,....., ... !~,.{, .."'(' l"nra'lC'''' .llVJ.L vlJ.VlJ.vJ,..I.,) L.I.....-\"'/L. '";:] ........

charging time required.

nf fl,p;.,.. '--~ \.1", ....... ..1.1.

1i""'1't~r1 .1 ........ J.I.

\..- ......

r'1nrrp.

)..(~J.lb'-'

u

nd

'tJA't~fO't'"tn~nrp 1-'·--.i~ ....... ":"__ J..A_'':,

. . . .,.,rl

1..;,'-.-,

t-r n

.. " - - '

v h'1lt0r ".'--".,J ,

:J

battery

H2CUraJDl:S liCe]

oppe,

1 lOJ?P

Electric motor! generator

Power electronics - controller -

Figure 3.9. Components of an electric drive

Electric vehicles are therefore likely merely to complement conventional vehicles, and therefore account for at the most 10% of the total number of passenger cars [3.8].

3.2.3 Hybrid Drive Hybrid drives are drives that have at least two different prime movers and energy accumulators [3.2]. Possible energy accumulators are

o o o

chemical energy accumulator: electrical energy accumulator: mechanical energy accumulator:

- conventional fuel tank, - battery, - flywheel, - hydraulic accumulator. Mechanical energy accumulators have up to now been used principally to assist moving off. They are much used in vehicles with a high proportion of stop and go driving, for example city buses, where recuperated braking energy is used. But there are designs under development, in which flywheels running in a vacuum on magnetic bearings act as the energy accumulator for emission-free operation [3.9]. The mechanical energy of the flywheel or gyro store is converted into electrical energy and feeds the electric motor. In hybrid drives, numerous drive combinations are possible. But they all have the problem of additional weight. Internal Combustion Engine + Electric Drive The combination of combustion engine plus electric drive gives better range and availability than a vehicle with electric drive only. In conurbations, driving can be emissionfree. Vehicles with hybrid drive are less energy-efficient than vehicles with just one type of drive, because of their additional weight. Regenerative braking is used to offset the increased energy consumption resulting from higher driving resistance. In order to keep the vehicle weight within bounds, "savings" have to be made with both types of drive; compromises are necessary. It has to be decideq whether there should be a preferred type of drive, and which it should be. With hybrid drives having both a combustion engine and an electric motor, a distinction is made between

o

serial hybrid power train (Figure 3.10): - no mechanical coupling of combustion engine and wheels, - no mechanical transmission required, - two electric units are needed (generator + generator/electric motor),

.:::-x~i~':~1' d,~''J:'\:;;;.:~:' ,:?~.rv:~'

"i,"'.,

'··;·:,,-i·'

3.3

Power Output, Combustion Engine Characteristic

ENERGY STORE

Tank

Battery

DRIVE LINE

Generator t--~

CONTROL

1.----_-1 Control

Chemical energy

71

o

Electric motor/ generator

Axle

unit t - - - - - - - - '

r\.

Mechanical energy

Electrical L,/ energy

o

Mechanical energy

Figure 3.10. Serial hybrid power train. Source: HOHN

o parallel hybrid power train (Figure 3. I I): - both dri ve sources can be combined, mechanical transmission is necessary, only one electric machine required, the battery does not need to be recharged from the mains. When the internal combustion engine is being used, it can run at optimum efficiency. If the specific power output of the engine is higher than that required to overcome the driving resistance, the excess power can be used to charge the battery. But note that the conversion losses during charging have to be taken into account.

-ru-

Tank

Combustion ' - - _ / engine Control unit

Mechanical

Axle

gearbox

Electric motor/ generator

Battery

Figure 3.11. Parallel hybrid power train. Source: HOHN

3.3 Power Output, Combustion Engine Characteristic Internal combustion engines based on the spark ignition and diesel principle will retain their dominant position in automotive engineering for the foreseeable future. ~park ignition engines are usually used in passenger cars. The key features are the hIgh power/

I

JI-

. ht

'V~,lg,

.

rl

f

rl

r~tl(), gOOt.t 1!~~~-,-'~~rn'""!3:1C~ :l~~_I. J

1... C)'\',' C()m0~1S h j • • 'T'1 0 r, rl" n " ' j " r, 0 n 0 " 0 11, 0 .. :C)n l1C1~::::. -" ",;.\-. . r1;,' v..&.,,-)~ ~ l/q.u.l"~b'--'~-' ~u..,--, LJ.J.v

quality of fuel required and the high part-load consumption.

..,---

...

.•

o

The economy of the diesel engine lis based on its low consumption, especially in the part load range, its low maintenance requirement (no ignition system), the low fuel quality J.~ljuiH;d aUel its gooJ gaseous emisslOIls ratings.! he dIsadvantages are the level 01 particulate emissions, noise, irregular running, the lower engine speed spread (11 max/ l1 m in), the low power output per litre, and the resultant greater weight and higher price. The higher capital cost means that it only becomes economical in vehicles which do a high mileage. Almost all commercial vehicles use diesel engines.

3.3.1 TorquelEngine Speed Characteristic There are two typical characteristic curves to describe the engine characteristic of combustion engines. One is the torque/engine speed curve at full load (100% accelerator pedal position) and the other is the corresponding full-load power curve (engine characteristic). Figure 3.12 shows the map of an internal combustion engine and the characteristic points of the full load characteristic curve. The maximum braking torque (0% accelerator pedal position) increases almost linearly with engine speed to a maximum of approx. 30% of the nominal torque Tn . Various measures are used to facilitate comparison of different engines. Key variables are torque increase (torque elasticity) Tmax

(3.24)

T=--

Tn and the engine speed ratio (engine speed elasticity) (3.25) Characteristic points

Power P(n) (engine characteristic)

I I

------~--------I

=

I

P (Tmax )

I

I I

t

I I I

I

T max

I I

.... Q)

I

-Torqu-e Yen)-

Tn

(full-load characteristic curve) I

I I

---+---i-----j 0,.,

Braking torque (thrust characteristic curve)

c c

Engine power at maximum torque

T max Maximum engine torque

T(P max ) = Tn Engine torque at maximum power = Nominal torque

n (Pmax )

=nn

Engine speed at maximum power Nominal speed

=

n (Tmax ) nn Engine speed n

(l)

ill

I

I I I

nmin

~c.. '0,

I I

o JL

=

P max Pn Maximum engine power Nominal power P (T max)

Pmax

,

n max ...

n (Tmax) Engine speed at maximum torque

Figure 3.12. Characteristic curves of an internal combustion engine

3.3

73

Power Output, Combustion Engine Characteristic

An engine is considered to have greater elasticity the greater the product r v is. This is apparent in the form of better engine power at low and medium engine speeds, which in turn means less frequent gear changing. Figure 3.13 gives examples of the characteristic curves for various passenger car and commercial vehicle engines. Different engine characteristics can be achieved by varying the engine design. A distinction is made in principle between three typical characteristics (Figure 3.14). a) Passenger car, Tprofile: Buffalo shaped

d) Commercial vehicle, Tprofile: flat

280

140

1200

Nm

kW

Nm

100 f- 100 ill

e) Commercial vehicle, Tprofile: Buffalo shaped

280

~

160 w ill 1400

>- 20

1200

0

1000

I

7

400 l.-

"-

/

"-

kW 320 CD ~

280 8240 ~ '5> 200 ill I-

Diesel engine 8 cyl., 141 l160 Turbo + Intercooler -, 120 I I

1000 2000 3000 4000 1/mm 6000

700 1000 1300 1600 1/min 2200

Engine speed

Engine speed

Figure 3. J3. Engine diagrams of various passenger cars a)-c), and commercial vehicles

d_)-_f)_ _- , - - - - - - - - - - - - - -

r /1·'·J .,

."

•

-I

1

.

Q.

Q)

//

.S 0) c:

Tprofile:

-

4I-----.~-l:--I => P profile:

LU

Rising Rising rapidly "Constant power"

Flat

Rising * _ . Buffalo shaped **

Engine speed n

Engine speed n

Figure 3.14. Typical engine profiles. *) Typical of 4-valve engines. **) Typical of commercial vehicle turbo diesel engines

3.3.2 Engine Spread, Throttle Map The spread of the engine is an important variable influencing the interaction of the internal combustion engine with the transmission (Chapter 5). In its function as a speed and torque converter, the transmission has a range of ratios: the overall gear ratio (Section 4.3.1). It is defined as the quotient of the maximum and minimum transmission ratio. The engine spread refers to an engine's speed and torque range. Vehicles with powerful engines accordingly have a large torque spread. Diesel engines have a lower maximum speed than spark ignition engines and accordingly have a small engine speed spread. Figure 3.15 shows the characteristic diagram of two passenger car engines. The engine shown on the left is a 2.0 litre spark ignition engine with a power rating of 111 kW; the one shown on the right is a 2.5 litre turbo diesel engine with intercooler. rated at 105 kW. 300 - . - - - - - - - , . - - - - r - - - - - - , 200 -.-------,-----y-----..., Nm Accelerator pedal 100% position 100% Nm ~~~=:::~ 60% 250.L~~-&:~~~-l- 150 -1----::;oo~

50%

=_..e:--1'~-T- 40%

(1) :;:j

30%

E'

oS 100 -4-~~-\---~~--j Q)

c:

20%

~

50 l-J~-~:----r---I

'en

~ 200 -J-_ _--#:>...l---=~~"d_~--::::.9=0..:..:%'_l :;:j

e-

.8 150 ..l----,~~~~~;;;:1 ~

70%

"g> 100 .L---If--\-P"""'~~--'~-6-00~Yo UJ

10%

/

/

'

0

2000 1/min 6000 o 2500 Engine speed n a) Engine speed n elerator pedal position . F 3 15 Passenger car engin~ p~~forman~eer:~i~'t~~ kW, 6 cylinder, 24 valves~ 19ure · ') a) 20 litre spark-lgmtlOn engl n d 105 kW 6 cylinder ' (throttle map . d: 1engine with intercoo1er an b) 2.5 litre turbo- lese

11 . min

7500

0 0 b)

3.3

Power Output, Combustion Engine Characteristic

75

. The diesel engine has a smaller speed spread than the spark ignition engine, but a greater torque spread. The transmission ratios have to be selected to accommodate this. The spread of the engine and the overall gear ratio (in combination with the graduations between gear ratios) are the main factors determining the functional characteristics of the vehicle. In the displayed engine maps the driver uses the accelerator pedal to indicate the power desired from the engine. When the accelerator pedal is fully depressed (100%) this corresponds to the engine full load curve, and when the accelerator pedal is not depressed (0%), to the thrust characteristic curve. Figure 3.15 shows the lines for the same accelerator pedal position for the two engines. The almost equidistant pattern is typical of diesel engines. The term "throttle valve angle" is often used instead of "relative accelerator pedal position". A throttle valve angle of 90° then corresponds to the engine full load line. Diesel engines do not have a throttle valve for preparing the mixture, so the term "relative accelerator pedal position" or "control rack travel" is used.

3.3.3 Consumption Map The fuel consumption of a stationary internal combustion engine can be represented as a function of engine speed and torque. A consumption characteristic diagram of this type is shown in Figure 3.16; the absolute consumption habs is shown in glh. It increases rapidly with the engine output. If the specific consumption he is shown i~ glk,wh, the t~rm "onion diagram" is used (Figure 3.17). There is a minimum consumptIOn hne he, min J~st below the full-load characteristic curve in the lower engine speed range. The precIse position depends on the engine. In the case of spark-ignition p~sseng~r ca~ engines. the minimum consumption is around 250 g/kWh, and with commercIal vehIcle dIesel engmes '" it is around 190 g/kWh. . In the ennine map the effective average pressure Pme m the cylInder IS often plotted instead of thebengine torque. The following relationship applies

Pme =

TM 2 Jr V . H

with

i=

l

2 , number of strokes

(3.26)

(1 bar = 105 N/m 2 ), where VH is the total swept volume in i = 0.5.

01 3 .

In four-stroke engines

Full-load characteristic curve Conversion: Absolute fuel consumption babs into specific fuel consumption be

IQ)

::l

0.....

o

Q)

be

c '51 c

[ g] = kWh

30 1000

1t

T {Nm] n

[---tl min :J

[lL]

babs h

w

Fuel density Pfuel : Petrol: 0.73-0.78 kg/I o 1-----:-~~~""7=:-~:""'":"::~=--........... Thrl..l~tra~ Diesel: 0.81-0.86 kg/I Thrust characteristic curve --....... Engine speed n

JL...

.

F_i_gU_r_e_3.16. Consumption map of a spark ignition engine. Absolute consumption babs In g/h

.

200 T--:--r:-:--I~:---r-~--r-~-----,-----. ~

~\

\

\\

o

..

~~ \

\

!:/

T( P in kif'!) •

111

IQ)

80

::J

e0

...... 100 Q)

c:

60

'8>

c: ill

\ 50

"'-

........

--

-- --

--

.

500

....

----. ....

....

..

-

..

--_ .. --

20

....

O--t----:--+--.,.---+---r---~-,.__-+_-~-!__-,.__~

500

1500

2500

40

3500 Engine speed

4500

1/min

10 5

6500

n

Figure 3.17. Engine map ("onion diagram"), specific fuel consumption be in g/kWh. Consumption map of a 2.0 litre spark ignition engine with II J kW, 6 cylinders, 24 valves

Like the engine characteristic the consumption map is an important basis for matching engine, transmission and vehicle. The transmission exploits the fuel-efficient areas of the engine performance map. Figure 3.17 shows the contour lines of constant specific fuel consumption be (onion curves) as well as the torque/engine speed curves at a constant engine output (demand power hyperbola) T(P = const). In this way the same engine output can be achieved at different torque/engine speed values - points 1 and 2 in the engine map - and thus also with different levels of fuel consumption. A minimum fuel consumption point can be found on any power hyperbola. The curve passing through these points is the minimum fuel consumption curve.

4 power Conversion: Selecting the Ratios

The transmission puts the engine to work

Chapter 3 dealt with the relationship between the power available from the engine and the power requirement arising from driving resistance. The torque/speed profile of the internal combustion engine is not suited to use in motor vehicles (see also Section 2.3.2 "Why do Vehicles Need Gearboxes? "). Output converters are needed for the final output to approximate as closely as possible to the ideal engine characteristic with P max = const over the entire engine speed range. Clutches serve to adapt engine speed, transmissions serve to adapt both speed and torque. The conversion ratio is determined by theoretical and practical engineering constraints, which depend very much on the application. The basic design of the transmission involves first determining the maximum and minimum ratio, i.e. the "overall gear ratio" of the transmission, and then selecting the intermediate ratios. Chapters 4 and 5 deal with the selection of these key features. They are the basis for the calculation, engineering and design of components (Figure 4.1).

Chapter:

3

Power Conversion: Selecting the Ratios

Optimisation involving: - Road trials - Test bed trials Computer simulation

.....

o..UJ UJCI) 0«

ZI 00..

Mediating the Power Flow

Matching Engine and Transmission

o

FINALISING THE TRANSMISSION PARAMETERS

DESIGN AND CONFIGURATION

..... Z UJ

~UJ

o..CI)

0« -.II UJQ..

> UJ o

l I

r!l 7

Gearwheels

9 Shafts

Synchronisers

10

Hydrodynamic cl:...;tchcs and converters

11

Bearings, casin~s,

etc.

Figure 4.1. The ratios selected are the key features of the transmission, and thus form the basis for subsequent development work

,

4.1 Power Train In vehicles with internal combustion engines, the output conversion between the engine and the drive wheels is achieved by the combined action of the assemblies of the power train. Figure 4.2 shows the hierarchical structure of the various ratios in the power train, starting from the total power-train ratio iA. The total ratio of the power train is derived from the ratio is of the moving-off element, the ratio iG of the transmission and the final ratio iE. (4.1 )

c

o :..;::; en

::J

.0

E

o

u

Cl3 Q) c~ .c Q)

.....

~

0>

C

Q)

I I I is I I I.e I .8 I u

iA

I

.iG, V

iE

I

I

I

iG, H

iG, N

I

iE , V

i E, A

C

"'0

::J

0

.'!: C ::J

Q) .....C .....C .....Q) ::J .'!:

'(j)

.~

0~

LL

I

0

Q..

E en

.'!:

"'0

.!:Q

I

I~

[Ioooooo~

iG

c

'(ij ~

E en

C

........ n:l

I

~

n:l Q)

.....CQ)

::J

0

a: E

c

Q) ~

::J

'+-

Q)

C

0> C

a1 ~

a1 ~

''';::;

n:l .'!:

'+-

n:l

en

en

x

0

~ .0

a1

.e.....

X

Q)

c

.~

Q) ~

C

~

~

Q)

c :=

::J

"'0

I

I~ ~I

I c!J

z /

/; ~I

ENGINE

I~ ~I

SELECTOR GEARBOX

FINAL DRIVE

Figure 4.2. Hierarchical structure of the power-train ratio iA using the example of a commercial vehicle with standard drive, i.e. front-mounted engine with rear-wheel drive

The ratio of output speed n2 to input speed n I of a power-train component is defined as speed conversion v, (4.2)

.

The torque conversion f.-i represents the ratio between the output torque T2 and the input torque T) of a power-train component, (4.3)

_.__._..,------....,-------'------------_.

4.2

Speed Converter for Moving Off

A ratio of i this case

;t:.

79

1.0 should only arise when there is both speed and torque conversion. In

if Jl > 1.0 .

(4.4)

Master clutches convert only rotational speed, i.e. is = 1.0. Torque converters convert both rotational speed and torque, is;;::: 1.0. Torque converters are discussed in Chapter 10. The discussion below is based on a dry clutch as the standard moving-off element. The transmission ratio io constantly adapts the traction available from the engine in steps - or rather in an infinite number of steps - to the traction hyperbola for P max = constant (see also Figure 2.22). Range transmissions are fitted on the input (iv) or output (iN) side to increase the number of speeds where the vehicle needs a broad overall gear ratio, such as commercial vehicles and off-road vehicles. The balance between performance and fuel consumption is achieved by adjusting the final ratio iE, especially important for commercial vehicles.

4.2 Speed Converter for Moving Off Internal combustion engines have a minimum rotational speed. The speed difference between the lowest engine operating speed and the stationary transmission input shaft has to be bridged by a speed converter. Frictional engaged clutches are always used as movingoff elements. Figure 4.3 shows a systematic classification of master clutches [4.1]. The main moving-off elements to have established themselves in motor vehicles are:

o

o

the dry clutch where is = 1.0 is standard for manual transmissions, the torque converter where is ;;::: 1.0 is standard for conventional fully automatic transmissions.

Less common is the magnetic powder clutch in which a magnetisable powder transmits the power by frictional engagement. Magnetic powder clutches, and also wet multi-disc clutches, are used as automatic clutches in continuously variable transmissions, for example. In dry clutches the pressure force is produced by a spring. A distinction is made between coil spring clutches and membrane spring clutches depending on the type of spring used. A distinction is also made between clutches activated by pulling and by pushing. Shiftable, master controlled, frictionally engaged I I

I

Frictional

Hydrodynamic

,

.......

Hydrostatic

Electrodynamic

Electrostatic

Magnetic

I

\ \

GJ

[£]

I

-I

I Magnetic powder

Clutch

"Converter"

Figure 4.3 Systematic classification of master clutches on the basis of their characteristics

,

The characteristic features of a speed converter are (Figure 4.4):

o o o

the output torque T2 is equal to the input torque TI: the output speed n2 is less than or equal to the input speed n I: the input power PI is reduced by the power loss Py:

T2= T], n2 ::; n] ,

P2 = PI-py.

SPE.ED CONVERTER

Figure 4.4. Speed converter input and output values

Figure 4.5 shows an idealised clutch operation sequence when moving off. The input and output speeds converge in the course of the clutch operation sequence. Some of the input power is converted into waste heat during the continuous slip phase. The efficiency of the clutch 'le is given by Equations 4.2 and 4,3 and is (4.5a) if T2

= TI

, i. e. J1 = 1, then

(4.5b)

Engine speed n1,O

c::

Reduction of engine speed

-1----"'---..,.

I I

nM, min

nM, min

"C

I I I

CD CD

a.

(f)

Speed of gearbox input shaft Slipping lime

n2,O

Disengaged

t:i~

Engaging phase

Engaged

Figure 4.5. Idealised moving-off sequence with a friction clutch

Time t

4.3

Total Ratio and Overall Gear Ratio

81

The slip S is defined as the ratio of the difference between the input and output speeds to the input speed (4.6) Equations 4.5 and 4.6 give the following relationship between efficiency, slip and speed ratio S

=1-

'lc

= 1-

Vc .

(4.7)

The master clutch must be so designed that it both transmits the maximum output torque with sufficient reliability, and tolerates the thermal stress arising in repetitive "stop and go" use [4.2].

4.3 Total Ratio and Overall Gear Ratio The power train has to offer ratios between engine speed and road wheel speed enabling the vehicle to:

o o o

move off under difficult conditions, reach the required maximum speed, and operate in the fuel-efficient ranges of the engine performance map.

The maximum ratio required iA, max is fixed by the first condition. The second condition gives the maximum road speed ratio iA(V max , th)' The smallest power-train ratio iA, min is given by the third condition. Figure 4.6 shows the speed spread of a transmission in a diagram of velocity against engine speed. The engine speed range (primary side) is "spread" by the transmission to the speed range of the secondary side. The operating range extends between the ratio boundaries.

Gearbox speed spread (seconda ry side)

Engine speed

min

... Figure 4.6.

n

Engine speed spread (primary side)

Velocity/p:ngine-~pecddi2.grar:1,

max -----1~

eJVcrall gcar ratio

Increasing legal constraints and traffic density are reducing the importance of maximum ~peeds of passenger cars. By the same token, acceleration performance is gaining in Importance.

L--.--,--------,---"!""'---------

A wide overall gear ratio is particularly important for heavy passenger cars with powerful engines and a low drag coefficients [4.3]. They need:

o o

a high stall torque ratio iA, max for moving off and accelerating, a low minimum ratio iA, min for low engine speeds at high road speeds to reduce fuel consumption.

4.3.1 Overall Gear Ratio The overall gear ratio of the transmission, often referred to as the range of ratios, is the ratio between the largest and smallest ratio . IG, max IG, tot = . IG,min

with the gears n = 1 up to z .

(4.8)

The overall gear ratio depends on:

o o o

the specific power output of the vehicle (P max / (mF + mpayload) in kW/t), the overall gear ratio of the engine, (see Section 3.3.2), and the intended use. Vehicles with a low specific power output, such as commercial vehicles, need a larger overall gear ratio. The same applies for vehicles with diesel engines, which have a small engine speed spread. Reference values for overall gear ratios of various vehicles are shown in Figure 4.7. Truck> 16 t liiii· . .··.i

i"

I'

iii

!.-

i·.

I

I Truck < 16 t I

Bus, long-distance

Pass. car: ace. & consumpt. opt. I,

.i·Fuh.Jre ?> .• »...>

Pass. car, diesel engine; transporter

I

Passenger car, spark ignition engine Passenger car: automatic

,~

o

I

2

4

Bus, urban: automatic I

I

I

6 8 10 12 Ratio spread iG , tot

14

16

18

Figure 4.7. Reference values for overall gear ratios for various types of vehicle. In the case of automatic transmissions, the conversion of the torque converter (Pmax ~ 2-3) has to be added

For passenger cars in particular it is necessary to consider that: o However great the overall gear ratio is, the transmission can only move the operating point on the demand power hyperbola (see also Figure 3.17). The most fuel-efficient range cannot be exploited by a passenger car with a powerful engine travelling on the level at moderate speeds since there is "insufficient power required". The engine and all power-train components have to fit together: power-train matching, see Chapter 5. Overdrive gears (iG < 1.0) result in reduced gear efficiency.

o

4.3

Total Ratio and Overall Gear Ratio

83

4.3.2 Selecting the Largest Power-Train Ratio The greatest traction requirement must be known to determine the ratio of the gear with the largest torque multiplication. The adhesion limit - i.e. the maximum force that can be transmitted between the tyres and the road - is a physical limit and must be taken into account when establishing the traction Fz, A at the road wheels (see Equation 3.9) F z, A :s; FZ , max = J1H R . Table 3.2 gives static friction coefficients J1H for certain operational conditions. Air resistance may be ignored at the speeds anticipated in the lowest gear. At the drive wheels a balance must be struck between the maximum requirements of acceleration, gradient, road surface, payload and trailer load: Maximum traction available Fz, A

TM, max iA, max

'/tot

= Maximum traction required Fz, B

_1_ = mF g (fR cos aS t + sin as t ) + mF A a rdyn

(4.9)

The largest ratio iA, max, often called the stall torque ratio, depends mainly on the specific power rating (kW/t) of the vehicle. Two extreme conditions may be considered:

o

The maximum gradient that can be climbed at an acceleration of a = 0 m/s 2 . Climbing performance, Section 5.2.2,

o

The maximum acceleration on the level. Acceleration performance, Section 5.2.3.

The stall torque ratio for passenger cars and commercial vehicles designed for maximum gradability is, from Equation 4.9:

.

rdyn mF g (fR cos aS t + sin as t )

lA, max =

(4.10)

TM, max '!tot

The dynamic wheel radius rdyn of most common tyre sizes is shown in Table 3.3. Reference values for rdyn are: ~ 0.3 m for passenger cars and ~ 0.5 m for commercial vehicles. Reference values for the rolling resistance coefficientjR are shown in Table 3.1. A climbing performance of q'max greater than 50% is normally required for an unladen passenger car. This ensures that a trailer can be towed and steep ramps mounted with ease. Acceleration performance depends not only on the stall torque ratio, but also to a significant degree on how closely the gears approximate to the traction hyperbola. The acceleration performance required depends very much on the brand image of the. vehicle. . The largest ratio in commercial veh::cles is often dictated by the vehicle's intended use. For example, building site vehicles and road sweepers have gears for extremely slow movement \v craw \)' Using the kinematic relationship (4.11)

.the crawler gear in a commercial vehicle is given by

Jl__---,--_,-..,..----,-

'

3.6 ~ nM min [_1_.] rdyn [m] 30

mm

'

(4.12)

where nM, min i~ in 1/min, rJyn is in m and known as crawler gears.

Vcrawl

in km/h. These very h1gh-ratio

gear;)

ale

4.3.3 Selecting the Smallest Power-Train Ratio Assuming there is no slip in the power transmission from wheel to road and that the (nominal) maximum speed is reached at maximum engine speed, then the smallest power-train ratio is given by

3.6 ~ nM max [_1_.] rdyn [m] 30' mm lA, min =

[

u max where

nM, max

k:

(4.13)

]

is in l/min, rdyn in m and U max in krnJh.

Commercial Vehicles The limiting factors of legal speed restrictions and diesel engine cut-off speed mean that the maximum speed will often be a design parameter when developing commercial vehicle power trains. The design ranges for commercial vehicles in Europe arising from the maximum permissible speed Umax are shown in Figure 4.8. 140

Maximum permissible speed

kmlh ::::. >-

....

'0

80

Q)

~

C/)

~

0

>

Bus with 100 kmlh motorway limit

100

60

~ 0

40

C\l C\l

A

20

~

(.)

o

0

'V ,...

Q)

t:: 0

....0

+= (.)

::J

::J

a.

~

(.)

::J

::J

l-

I-

t-

::J

.c

t-

(.)

t:: C/)

(\1

::J

+-0

.c

.... l0-

C/)~

t:: (.)

o

Truck motorway and trunk roads

::J

O,::;

t::

(\1

.c t:::>

C/)

U I

C)

t::

o

.....J

....L-..h-_..L.-I......I~--L--'--'-_....a.-~..L.-.................- - L _ - ' -.....

Figure 4.8. Design speeds for determining iA, min for commercial vehicle power trains. The maximum speed data relates to Germany

Passenger Cars There are various factors to be considered in selecting the smallest ratio. One factor is the high proportion of duty time spent in the highest gear, which can be more than 80% in the case of passenger cars. Depending on the type of design selected, a distinction is made between:

4.3

Total Ratio and Overall Gear Ratio

85

1/ U max - optimum design: iA, min = iA(U max , th), 2/ overrevving design,

.

3/ underrevving design.

1/ Vmax - Optimum Design In order to convert the maximum engine power installed in the vehicle into maximum performance, the required power curve Pz, B must pass through the point of maximum engine power available Pz, Amax (= P n ) [4.4]. This is called the design point A, Figure 4.9. It represents the maximum speed u max , th theoretically available (q' = 0%; no wind). The acceleration reserve and fuel consumption in top gear are also important factors· in the case of passenger car transmissions. The excess power available Pz, Ex is a measure of acceleration reserve, and the engine speed nM serves as a measure for fuel consumption (Figure 4.9). 2/ Overrevving Design The power available and the power required intersect in the declining section of the power supply curve Pz, A as shown in Figure 4.9, Point B. The speed Umax 2 which can be achieved at this point with an overrevving design is less than u max , tho The power-train ratio iA2, min is greater than iAI, min = iA(U max , th). This is achieved by increasing the ratio of the highest gear iz or the final ratio iE, Since the engine speed is then higher for a given road speed, the operating point moves into the range of higher fuel consumption on the engine map. The high level of excess power Pz, Ex2 makes this arrangement preferable for sports designs. 110 -,----...----.....,.---,------r--,----r---;------, PZ, Amax = Pmax 11 tot

kW

90 -+----+--~~I_+____::;Itf1C_-t------J'_t:_1~--+__----t------'____j

a......

Q)

3: o a..

PZ , A at:

./.

iA2 , min = 3.7

"

70 ~----:--~+-----l~-;i .. = 3.25,," A1

I "

,min

iA' 3 ,min .

...

50 " 110

."

=2.7 130

170

150

190

210

km/h

6034

1/min

5014

5~91

250

Velocity v

3161 3599 2626

3736 4253

4310 4907

3i04

3501

14885 15562 L4058

5460 6216 4S3S

5968

Engine speed n M

Figure 4.9. Selecting the ratio in top gear for passenger cars. The excess and associated engine speeds are designed for a speed of 170 km/h. Design: 11 at v max , th; 2/ overrevving; 3/ underrevving

•

~,.~.

,-,vavCfJtUfl.

IJeU::Cllng

rne l

Ratios:

60 -;----;-~"'f_t--_+_::.,;c_-I__t_-~.-r=:--+--l

: 3.44

1.35

5th gear

i 5 : 2.54

1.37

40 -+----::r-~~'::::E7"j-~%~;;;;;;;;;;;;~ I 4th gear ~)0111::;;"-'-~~ 3rd gear

i 6 : 1.86

1.38

i 7 : 1.35

1.35

1st gear 201~~g~~~~~~~;~;;;9l2nd gear

i 8 : 1.0

i4

O-+--r--r--+--r--+--r--r-r--r--I--'--r--r--r--r--+--+--r--r-~

500

1000

n min

2000 1500 Engine speed n n

1/min 2500

max Figure 4.11. Velocity/engine-speed diagram of a bus with 8-speed 2-range transmission. Maximum road speed in the diesel engine governed range. The maximum speed is reached in 8th gear in the "governed range" of the diesel engine (see also Figure 5.6). The saw profile diagram shows the earliest upshift possible without stalling the engine, and the earliest downshift possible without exceeding the maximum engine speed.

4.4.2 Geometrical Gear Steps In the geometric design the gear step ffJ between the individual gears always has the same theoretical value ffJth =

Z-~iG, tot

The ratios of the individual gears n . _.

In - I z

(z-n) ffJth •

(4.15)

.

= 1 to z is then given by (4.16)

~4.4

Selecting the Intermediate Gears

89

';!npractice the ge~ step will vary sligh~ly from lpth (Fig~lfe 4.11). The .approximation to the effective tractIOn hyperbola Fz, Ae IS equally good III all gears (FIgure 4.12a). The . difference in maximum speed between the gears consequently increases with each shift to a higher gear. Geometrical gear steps are most common in commercial vehicle gearboxes; the lower specific power output means all the gear steps are of equal significance. Range transmissions (Figure 4.11) have to be geometrically stepped to make all ratio steps the same size, preventing individual gears from overlapping (see also Section 6.7.2 "Multi-Range Transmissions").

,i

4.4.3 Progressive Gear Steps Progressive gear steps are used for passenger car transmissions. The higher the gear, the smaller the gear step. Figure 4.12b shows the progressive transmission stepping in the traction diagram and the velocity/engine-speed diagram (saw profile diagram). 8 _..._---rr---,..---.....-----.----, i

1 = 4.14

1

2

· Ip

= "if =

· Ip

n2 1 ~ = n;= 1-i = 1+ ~;

· Ip

=

3

~:Ci)

caca> Q).Q> c rJ).t:: ca.~

a..ErJ)

"0

2 2

1

3

c

f)

· Ip

2

C

e)

Planetary step ratio

a> ca..c x a> .... ca C).§

0>

Spider Ring gear

--

~~rJ) ~'-

d) 1

-"0

Output Frame

....e:!

1

3

2

.. Ip

n1

2

Zl zl

=-T 3

1 .

-IS

=1 + z3 Z1

S

n2 n 3

-L _1_ = 1-.,1 = 1 + ~ 'S

z3

1 Z = (1=1--. =1 +-1 n3 2

IS

z3

This table does not list the three trivial states of motion in which the transmission rotates as a block. Furthermore, not all transmission ratios are suitable for use in motor vehicles. If none of the parts in a planetary gear-set is in a fixed position, then it is referred to . as a differential drive, or a summarising gearbox, transfer gearbox or differential gear. If several planetary gear-sets are linked together, the result is a so-called coupled gear. This 80rt of gear makes it possible to achieve different ratios between input and output, depending on how the individual transmission components are linked together and which components are in a fixed position. The components are linked together by clutches, and the components are linked to the housing by brakes. The great variety of possible ratios available in transmissions with just one planetary gear-set is further substantially in.creased in coupled gears, but not all the ratios that can be derived from the transmission ·are relevant in motor vehicles. There are other important designs in addition to the simple planetary transmissions discussed here. You may wish to refer at this point to Section 6.6 and the relevant literature [6.8] to [6.10]. Traditional automatic transmissions with various gear ratios are made up of several individual planetary gear-sets. The ratios of the individual gear steps cannot be freely selected independently of each other, since the same gearwheels are used for several gear steps. A section from a Wilson transmission is shown in Figure 6.16 as an example of such a planetary transmission. The spider of the first planetary gear-set is connected to the ring ·gear of the second. The two gears are only shifted by the closing of the corresponding brake. Planetary coupling gears can also be power split, as shown in the illustration (Figure 6.16) above in the case of second gear. Reactive power can also occur in calculating the power values in the various paths. Reactive power can be envisaged as power, flowing in a circuit, which is not detectable from outside. But it stresses the components through which it flows, and impairs the overall efficiency of the transmission. Planetary transmissions can reach very low overall levels of efficiency, which in extreme cases can even · become negative. This represents transmission interlock, which in certain circumstances can be desirable if the transmission is not to be moveable from the output side.

8

8

'"

~

T

fL__

III 1st gear

2nd gear

Figure 6.16. Planetary coupling gear: Section from a Wilson commercial vehicle gearbox (British Leyland; four forward gears: four planetary gear-sets, four belt brakes, one clutch)

6.5 Fundamental Approaches for Part Functions, Evaluation In the concept phase of developing a transmission, basic approaches are established; see Figure 14.14 in Chapter 14 "Systematic Engineering Design". A large number of transmissions can be created by combining the individual approaches used for the main functions, as shown in the morphological table in Section 6.2 (Table 6.6). The number of viable alternatives is however significantly reduced when a technical/economic evaluation is carried out. This can be demonstrated using the example given in Table 6.8 for the main functions "Enable moving off' and "Change ratio". This is given as an example, and does not claim to be comprehensive. A complete evaluation of all proposed solutions for the main and ancillary functions of the transmission should be carried out after the concept phase. The design phase proper can begin when this evaluation has been completed. Table 6.8. Example of assessing approaches to the sub-functions "Enable moving off' and "Change ratio". O...not possible; l...very unfavourable; 2...unfavourable; 3...moderate; 4...favourable; 5...very favourable Approaches Friction Fluid clutch clutch

Function

Gearwheel

Pulley drive

Convert torque Vary slip

5

0 4

0 3

4

4

0

4 0

3

3

Efficiency

5

4

4

3

3

4

Service life Reliability Ease of use Space demand Price

4 2 4

3 2 3 4 4

4

4

4 3

4

5 5

3 3 2 3 2

2

3 4 2

2 3 2 2 1

Total points

30

21

24

23

27

21

4

Torque Hydrostatic converter gearbox

6.5

Fundamental Approaches for Part Functions, Evaluation

129

Table 6.8 suggests that the gear pair commends itself as by far the most cost-effective element for torque conversion. The disadvantage that this eliminates all but geared transmissions becomes a secondary consideration. Friction clutches are still the best available compromise for moving off and for speed synchronisation. The torque converter also has many advantages.

6.5.1 Reverse Gear There are numerous designs for implementing the ancillary function of reverse gear. Figure 6.17 shows six different variants. . a)

c)

b) Main shaft (MS)

=-tl--

-1

Countershaft (CS)

MS

n CS

--+H-f)

d)

l, \

\

-::::::::::=-....;::::::::::--- MS -4----+----

MS

LyLe n

CS ~1----4---

CS

L'R

MS

1!Sprocket

~ ~ chain - 4 - - - 1 - + - CS

Figure 6.17. Alternative reverse gear configurations. a) An axial sliding gear is inserted between each fixed wheel of the main shaft and the countershaft ; b) Shiftable shaft with two pinions between a reverse gearwheel of the main shaft and a forward gearwheel of the countershaft; c) The sliding gear is inserted between a fixed wheel of the countershaft and a toothed sliding sleeve of a synchroniser on the main shaft; d) Sliding shaft with two pinions between a forward gearwheel of the main shaft and a forward gearwheel of the countershaft; e) Reverse gear with intermediate pinion constantly engaged, shifting with sliding sleeve; 0 Reverse gear using gear chain, shifting with sliding sleeve

The required reversal of the direction of rotation of the gearbox output shaft is usually achieved by inserting an idler gear into the power flow. The general rule for toothed gearing is that increasing or reducing the number of ratio steps by one reverses the directIOn of rotation of the output shall. 1~0l ali the variaHls ~hlJwll ill Pigure; c.n are of equal significance in practice. The following highly simplified assessment in Table 6.9 is intended to highlight their strengths and weaknesses. Tf reverse fear noes not lise a gearwheel of a gear step of the forward gears, the cheaper spur gear toothing can be used for reverse, because of the relatively small proportion of time spent in reverse gear. The resultant increased noise level is acceptable.

J

Table 6.9. Advantages anddisactvantages of various types of reverse gear (ref. Figure 6.17); + advantage, - disadvantage

Solution .~ Evaluation criterion

.

a)

b)

c)

d)

Easy to synchronise Can be synchronised at rest

+

Saves components compared to a) No ratio or toothing constraints

+ +

+ + +

Sufficient shaft· clearance to accommodate the toothing

-

Reverse gear must be helical cut Practicability

-

-

f)

+

+

+

+

+

+

No axial space requirement

e)

-

6.6 Passenger Car Transmissions Passenger car transmissions are classified into the following main designs and types: a conventional 4-6 speed manual transmissions, a semi-automatic transmissions, a fully automatic transmissions: - conventional 3-speed to 5-speed automatic transmissions (consisting of torque converter and rear-mounted planetary gear), - 3-speed to 6-speed automatic countershaft transmissions, a mechanical continuously variable transmissions. The dominant design for passenger cars in the USA is the conventional automatic transmission. 75 to 80% of all vehicles are fitted with them. The same trend was predicted in Europe in the late '60s, but failed to materialise. Here the market share of automatic transmissions in passenger cars has remained virtually constant at around 15%. Multi-range transmissions, as discussed in Section 6.7.2 for commercial vehicles, are in principle also feasible for passenger cars.

6.6.1 Manual Passenger Car Transmissions Manual passenger car transmissions include all transmissions in which both the process of changing gear and the process of engaging the master clutch and moving off are carried out manually by the driver. They are all fitted with spur gears. Transmissions with dog clutch engagement are now practically unheard of in passenger cars. All transmissions are offered with synchromesh. Occasionally the reverse gear is still not fitted with synchromesh. Passenger car manual transmissions can be sub-divided into further categories (see also Section 6.3). This subdivision relates mainly to the design of the main gearbox itself,

I

6.6

Passenger Car Transmissions

131

not to any integral final drives, differentials and intermediate shafts needed to drive them. With this limitation, the following categories result: o single-stage countershaft transmissions with 4 to 6 gears and integral final drive, o two-stage, (coaxial) countershaft transmissions with 4 to 6 gears with and without integral final drive, o three-stage countershaft transmissions with 4 to 6 gears and integral final drive. Single-stagecountershaft transmissions are used in passenger cars in which the engine is located on the drive side, which is to say in rear-wheel drive vehicles with rear engines, or in front-wheel drive, vehicles with front engines. This applies to both the normal engine configurations - transverse and longitudinal; they also occur in "transaxle transmissions" (e.g. Porsche 944 with gearbox behind the axle) (Figure 6.2i). In the case of single-stage countershaft transmissions, the final drive is usually integrated into the gearbox housing. In the transmission diagrams used in this chapter, integral final drives, where present, and reverse gears of the various transmissions are represented by "faint" lines for the sake of completeness. It should be noted that in reverse gears the shafts of the idler gears are located in a different plane to the main shafts (compare also with Section 6.5). The location and size of the idler gears are intended only to give an impression of the fundamental design. Two-stage countershaft transmissions are used in passenger cars with standard drive. They normally contain no integral final drive components since they are generally flanged directly onto the front-mounted engine, and linked to the drive axle by a propeller shaft. One exception is two-stage transmissions mounted on the rear axle to give more even weight distribution with front-mounted engines (e.g. Porsche 928; transmission in front of the axle) (Figure 6.2h). Parts of the final drive are integrated in them. In three-stage transmissions, a part of the transmission is relocated to a third countershaft located elsewhere. These transmissions are used in passenger cars where space constraints dictate a very short overall gearbox length. The synchroniser packs are each allocated to one shift level, and serve mostly to shift two neighbouring gears. In each shift level there is usually a first and second gear, third and fourth gear, fifth and reverse gear, or alternatively fifth and sixth gear. There are also designs which use a separate clutch element for the fifth and reverse gear, which can be unsynchronised in reverse gear. 4th

3rd 2nd

1st

1J u

a)

L 1 •." . •.•• ,'

.'

b)

g g g ,.-F_i_u_r_e_6_'1_8_._a_)_Sin le-sta e 4-speed gearbox (VW); b) Single-stage (VW), production design Figure 12.5

5~speed gearbox

,

The example of a single-stage4-speed gearbox is the VW unit, as used for example in the VW Golf (6.18a). In this gearbox, the gear pair of the first gear is located directly beside a shaft bearing. The total number of gear pairs remains the same compared to a tvvo-stage 4-speed transmission, since although the gear pair of the input constant gear CG, (sometimes called head set), (Figure 6.19) is not required,one is needed for the fourth gear. Two-stage transmissions have a direct gear. The single-stage 5-speed transmission (Figure 6.18b) differs from the single-stage 4-speed transmission only in having an additional gear stage, which is "attached on to" the drive housing side opposite the input side. This does not require any design changes in the original gear unit. Numerous 5-speed gearboxes have been developed from existing 4-speed gearboxes. 4th

a)

b)

Figure 6.19. a) Two-stage 4-speed gearbox (Getrag); b) Two-stage 5-speed gearbox with direct 5th gear, "sports gearbox" (ZF), production design Figure 12.1

One example of the two-stage 4-speed gearbox is the Getrag gearbox in Figure 6.19a. In accordance with the design principle of placing gear pairs with high torque changes near bearings in order to minimise shaft deflection, the gear pair of the first gear is located on the gearbox output side. The fourth gear is the direct gear. In the 5-speed gearbox shown on the right in Figure 6.19, the fifth gear is the direct gear. Frequently the fifth gear is speed increasing (overdrive) and fourth gear is the direct gear. 6th 3rd

4th

CG 5th

3rd 6th w.L...1. ~: ?·.W~,:

a)

N..}..•..

b)

Figure 6.20. a) Two-stage 6-speed gearbox (Getrag), production design Figure 12.4; b) Single-stage 6-speed gearbox (Opel), production design Figure 12.6

I

6.6

Passenger Car Transmissions

133

There are manual gearboxes for passenger cars with up to six forward gears (Figure 6.20). In the two-stage countershaft transmission (Figure 6.20a), the gear step of the first and second gear are near a shaft bearing. It should also be borne in mind that such transmissions are used principally in high-performance passenger cars, and therefore have a high torque design. Figure 6.20b shows a single-stage countershaft transmission with final drive. The reverse gear is located on a lay shaft to save space for the front transverse configuration. The transmission is three-stage for the reverse gear. . The advantage of the three-stage passenger car gearbox design is its short overall l~ngth. (Figure 6.21). In fact this transmission is only three-stage in first and second gear, smce m these gears the power flow runs through the countershaft. In third and fourth g~ar~ the po",:er flow .goes directly from the input to the output shaJt, making the transmISSIOn functIOnally smgle-stage. The reverse .gear acts as a two-stage gear unit since the power flow runs directly from the countershaft to the output shaft.

Countershaft

Input shaft

Output shaft

Figure 6.21. Gearbox diagram: threestage 4-speed gearbox (Volvo)

6.6.2 Semi-Automatic Manual Passenger Car Transmissions The term "Semi-automatic transmission" relates to the two operations "Engaging the clutch/Moving off' and "Changing gear". One of these operations is automatic in semiautomatic transmissions (see Table 6.12 "Levels of automation of passenger car and commercial vehicle manualgearboxes").They may be broken down as follows:

o o

transmissions with automatic master/gearshift clutch, manual gear shift, transmissions with driver-controlled clutch, automatic gear-shifting process.

In the first variant the driver merely sets the desired gear by shifting the control lever, and drive take-up, moving off and changing gear is carried out automatically. In the second variant the driver preselects the gear or follows an automatic gear selection recommendation, activating the complete gear change with the clutch. In private passenger cars the first variant is principally used, whereas in commercial vehicles the second variant is common (see also Chapter 13 "EnginefTransmission Management"). Conventional manual gearboxes can be converted by retrofitting automatic controls for clutch operation or gear shifting. In the VW torque converter clutch transmission (1967) there is a mechanical gearshifting clutch mounted behind a torque converter (Figure 6.22). When the gearshift lever is activated, the gearshifting clutch is automatically disengaged, interrupting the power floV! to the manual gearbox. The gear can now be shifted manml11y. When the operation

l L,__

is complete, the gearshifting clutch engages again automatically. ~

-..,. .

_

, I Figure 6.22. Gearbox diagram: 3-speed torque converter clutch gearbox (VW 1967) The converter has three main functions to fulfil in this process: o Enable moving off in any gear. o Refine the coarse stepping (three forward gears) of the manual gearbox. o Damp the torsional vibration when engaging the gearshifting clutch. The main gearbox is a single-stage 3-speed gearbox developed from a 4-speed transmission by converting what was originally. first gear into a reverse gear. In practice this transmission concept had to contend with high fuel consumption. The reason for this was the constant power flow through the converter - there was no lock-up clutch - and the fact that with this transmission it was possible to move off in second or third gear. This design therefore never became popular in passenger cars. Transmissions with an automatic clutch are also to be found in Formula 1 racing cars. The driver activates a gearshift lever to manually control shifting up and down. In these transmissions only the gearshift clutch operation is automated, not the moving-off process. Developments in passenger cars aim at automating the mechanical gearshifting clutch and master clutch (such as the Mannesmann Sachs EKS "Electronic clutch system" or the LuK EKM "Electronic clutch management"). The systems are now in mass production (production design, Figure 12.16a). In these systems the driver only changes gear manually - the clutch is controlled by an automatic system, both when changing gear and when moving off.

6.6.3 Fully Automatic Passenger Car Transmissions The term "fully automatic transmission" is applied to geared transmissions in which the two part functions "moving off/engaging drive" and "changing gear" are carried out automatically in accordance with fixed or adaptive programmes. Fully automatic transmissions have some advantages compared to conventional manual transmissions, such as . 0 reduced driver stress, and consequently improved road safety and ride comfort, o faster shifting than the average driver, o "smarter" shifting than the average driver, and therefore lower fuel consumption. Gear shifting without any (noticeable) power interruption (powershift) is desirable in passenger cars to improve passenger comfort (see also Section 6.3.2 and Section 9.7). Countershaft-Type Automatic Transmissions Fully automatic passenger car gearboxes of one-, two- or three-stage countershaft design have the advantage of being very compact, allowing free choice of ratio, and comprising standard elements. Transmissions with high numbers of gears are simpler to build than with conventional automatic transmissions.

6.6

Passenger Car Transmissions

135

Well known examples are the Honda (Hondamatic, Figure 6.23a), the GM Saturn transmission and the W5A 180 transmission of Mercedes A-Class (production design, Figure 12.16b). In these transmissions a countershaft transmission is mounted after the converter. The conventional synchromesh units are replaced by multiplate clutch packages. The oil feed to the rotating multiplate clutches through the shafts is a soluble problem. Automatic wet and dry clutches can be used as the moving-off element as well as the converter. Another example of powershift passenger car transmission and (depending on the degree of automation) fully automatic countershaft-type passenger car transmission is the twin-clutch transmission (Figure 6.23b). Twin-clutch transmissions were developed for commercial vehicles and passenger cars. The Porsche electronically controlled twinclutch transmission (PDK = Porsche twin-clutch transmission) has been successfully used in racing passenger cars. The transmission input shaft is split into a solid shaft and a hollow shaft. There are two routes available for the power flow. In the example shown in Figure 6.23, the clutch C I is used for the second and fourth gears, the clutch C2 for the first and third gears. The gears in the part that is currently not active can be preselected. Gear changing is then carried out by shifting from the one clutch to the other.

1st

4th 2nd 3rd

4th 2nd 3rd 1st

1JL w

a)

I

b)

Figure 6.23. a) Hondamatic countershaft-type automatic gearbox (Honda); b) Twin-clutch transmission for installation on the transaxle principle (Porsche)

I

Conventional Automatic Transmission Fully automatic passenger car transmissions consisting of a to:que converter with a planetary type gearbox enabling shifting without power interruption are known as co~­ ventional automatic transmissions or just "automatic transmissions". Fully automatic . . transmissions are now predominantly of conventional design. Conventional automatic transmissions consist of the components lIsted above, wIth the power flow active in the particular gear step being defined within the planetary gearsets by clutches and brakes (Figure 6.24). . The Simpson planetary gear-set has manufacturing advantages. b~cause It has the same number of gearwheels in both transmission parts. Bot? transmIs~lOn parts.fU? on.a common wider sun gear. The design most commonly used 1ll.autom~tIc trans~IsSlOns IS the Ravigneaux planetary gear-set, (Figure 6.25). This makes It pOSSIble to achIeve up to four practically useable forward gears and one reverse gear.

JL'-~----r--..,...----_._-----------

f

Gear

1st

81

X

82

(X)

C1

X

C2 F

I

2nd 3rd R

X X

X X

X

X

Figure 6.24. Gearbox diagram of a Simpson planetary gear-set. C Clutch; B Brake; F Free-wheel; X Element engaged in power flow

The Ravigneaux set is a so-called reduced planetary gear. These are planetary transmissions in which the construction resources are "reduced" since parts of the individual simple planetary gears are the same size and can therefore be grouped together [6.8]. By using gear-sets of this type that can shift up to five gear steps, the overall dimensions of such planetary gears are relatively short. The selection of ratios is now restricted. Since the individual gearwheels are used for several gears, the resultant gear steps have to suffice. Since in the case of automatic transmissions the converter carries out part of the change in transmission ratio, they theoretically require fewer gear steps than the manual transmission. Some of the 5-speed automatic transmissions, which are becoming increasingly popular, use an additional planetary gear-set. Most of the space taken up by automatic transmissions is occupied by the multiplate clutch packs in the clutches and brakes required to shift the gears.

2

3

'---4

Figure 6.25. Ravigneaux planetary gear-set. 1 Common ring gear; 2 Narrow planetary gear; 3 Broad planetary gear; 4 Large sun gear; 5 Small sun gear [6.11]

6.6

Passenger Car Transmissions

137

There are two different types of brake as standard, the belt brake and the multi-disc brake. In the belt brake a metal belt runs once or twice around a brake drum, and brakes the drum by tightening the belt. This braking process is not easy to control, since the braking action is very rapid because of the self-reinforcing physical principle involved. In view of the increasing requirement for easy gear shifting, the multi-disc brake is becoming increasingly common. Although this takes more space than the belt brake, the shifting action is improved because of the finely controlled braking process. The multidisc brake is based on the same components as the multi-disc clutch, which serves to link the moving parts of the transmission together. See also Section 9.1 "Shifting Elements". The clutches and brakes discussed above for shifting the various gear steps are hydraulically controlled by hydraulic fluid. This fluid is supplied under pressure by a primary (engine driven) pump. The power it absorbs is no longer available to propel the vehicle, and thus represents an efficiency loss for the transmission as a whole. The effect of this pump on the overall level of efficiency is comparable to that of the torque converter. An overview of the losses in automatic transmissions is given by the highly simplified block diagram of a conventional automatic transmission in Figure 6.26.

Pv

~

Power loss

Clutches

Converter

Planetary gear Lock-up clutch

preS~~{iSedt Pv

I

Pump

I

Brakes

¥ I pres~~rised tl. i

.Pressurised oil

I- - - - - CONTROL UNIT

Figure 6.26. Block diagram and power losses ina (conventional) automatic transmission

The functioning of a 4-speed automatic transmission is considered in detail below based on the ZF 4 HP 14 automatic transmission (Figure 6.27). This transmission is designed for use in front-wheel drive passenger cars, which would become apparent in the transmission diagram only after the planetary gear. The components devoted to the final dri~e are not shown, since· they have no effect on the principle of operation of the automa~lc mechanism. The components involved in the particular gear step are shown by heaVIer lines. The ZF 4 HP 14 four-speed automatic transmission consists ofa torque converter with integral torsion damper T. To improve efficiency, the transmission has no torque converter lock-up clutch, but works with power split, (see Section 10.7). ~lease refer to Chapter 10 for a precise functional description of torque converters. There IS also ~ crescent design oil pump linked to the pump shaft of the converter (not show~ m t~e diagram) to provide the pressurised oil necessary to shift the ge~rs. The actual kl~ematlC transmission is a 4-speed Ravigneaux set The clutches are multIplate clutches shIfted by oi] prf:ssure.

tl . L--r-------r-_----,-

_

Figure 6.27. Gearbox diagram: 4-speed automatic ti'an~;mission

(ZF) in neu lia];

Trilok converter: P Pump, T Turbine, R Reactor TD Torsion damper F Freewheels B Brakes C Clutches TC

The brakes are of both designs, multi-disc brakes Bl and B3 and belt brake B2. The clutch linings and brake linings in automatic transmissions have an extremely long service life if correctly designed, since they run in oil, and are almost non-wearing. In 1st gear, the spiders of both planetary gears are retained by the freewheel F2, by means of which the planetary gear-set functions as a gear-set with fixed axles (Figure 6.28). The input power flows through the converter and the engaged clutch C3 to the large· sun gear of the Ravigneaux set, and back out of the planetary gear-set via the ring gear to the output. The effective ratio i = 2.41. In second gear, the small sun gear rests against the gearbox housing by means of the free wheel F} and the brake Bl. The input power flows through the converter and the engaged clutch C3 to the large sun gear, as in the first gear. The bar of the planet gear-set now rotates, and the planet gear-set functions as a reduced planetary coupled gear. The power flows again via the ring gear to the output, and the effective ratio is i = 1:37. The third gear is the most interesting from the point of view of its functioning. The transmission functions with power split, Le. a part of the drive power flows through the torsion damper T and the engaged clutch C2 into the planetary gear-set which functions as a differential drive. The second power split flows from the converter through the clutch C3 to the large sun gear of the planetary gear-set. Both power branches, or rotational speeds, "overlap" in the planetary gear-set, and are fed to the output at the ring gear. This operating status of the power split is not to be confused with that of a closed torque converter lock-up clutch TCe. In the torque converter lock-up clutch the impeller P and turbine wheel T of the converter are linked together, locking up the converter (Figure 6.29). The transmission ratio in third gear depends to a small degree on the slip in the converter, and is therefore not constant. The ratio in third gear thus varies between i = 1.0 ~ and 1.09. In 4th gear the converter runs without load, and power transmission to the planetary gear-set is purely mechanical through the torsion damper TD and the clutch C2. The Ravigneaux set functions as a simple planetary gear driven through its spider, and whose sun gear is supported at the housing through the brake B2. Output is through the ring gear. The ratio in 4th gear is i =0.74, constituting an overdrive.. In reverse gear the Ravigneaux set again works as a simple planetary gear reversing the direction of rotation. The power flows through the converter and the clutch Cl to the small sun gear. The spider is supported against the housing through the brake B3. The output is through the ring gear. The reverse gear ratio is i = -2.83. A further example is the 5 HP 18 five-speed automatic transmission for standard drive (Figure 6.29). In contrast to the 4 HP 14, this transmission has no power split. The converter can be removed from the power flow by means of a torque converter lock-up clutch CC. See also Figure 5.9 "Interaction of engine and converter". The fact that the torque converter lock-up clutch is never shown as engaged in the ZF 5 HP 18 does not indicate that this is never a function, but rather that it can be engaged optionally in each range.

Passenger Car Transmissions

139

1st gear LlLL

2nd gear

3rd gear

4th gear LlLL

Reverse gear

Figure 6.28. 4-speed automatic transmission (ZF); power flow in the gears

An important, if not the most important, assembly in an automatic transmission is the control unit. It is responsible for activating the brakes and clutches in the transmission. Their control has a direct influence on the "shifting qualitv" of the transmission as

perceived by the driver (see also Section 13.3 "Transmission Control").

cc

1st gear

4th gear

2nd gear

5th gear

......... L'-L-Io.;.:.:.;..

...L-I

3rd gear

Reverse gear

Figure 6.29. Gearbox diagram and power flow of a 5-speed automatic transmission

(ZF)

[6.12], production design Figure 12.17

In principle two types of control can be distinguished: o hydraulic control units, o electronic/hydraulic control units. In hydraulic control units the input information to be processed is converted in a purely mechanical manner into proportional oil pressures which activate the shift elements through the hydrostatic servos (in principle always pistons under pressure). The entire control algorithm is embodied in the design of the hydraulic control unit. The assembly is very complex, so reference is made at this point to the relevant literature [6.13]. This was the design used in the first automatic transmissions. It has now reached a very advanced stage of development, and normally functions without fault throughout the life of the transmission. Two factors have led to development of electronic transmission controls: the disadvantages ofhydraulic systems (such as the rigid control algorithm, embedded as it is in the hardware, and the fact that they cannot adapt to any mechanical wear), and the development of electronically managed engines. Electronic systems can adapt more easily to various engines or different operating conditions (adaptive gearshift programmes), and they can contribute to engine management (load reduction when shifting gear). They also have the general advantage of processing all available information, . through to controlling the shifting process, taking into account the vehicle as a complete system. The shift elements are still activated hydraulically even with electronic control units. The shifting profile of an automatic transmission is shown in simplified form in . Figure 6.30. The shift points are principally dependent on the vehicle speed and the load on the engine. The position and shape of the shifting characteristic curve are also adapted to current driving conditions in modern controls by evaluating other parameters such as lateral and longitudinal acceleration, or rate of change of the accelerator position. The driver can influence the shifting characteristics, e.g. the selection of an economy or performance driving style.

6.6

Passenger Car Transmissions

141

Kickdown Full load /------,..-1".--,.... Hysteresis~

(ij "0

ID

C.C

.... 0

o '';::

+-0'-

~

en

.... 0

IDc. Q5

Figure 6.30. Qualitative profile of the shifting characteristic curves of an automatic transmission

()

()

c< Idling

o.

Velocity v

---..

The hysteresis arising from the two different shifting characteristic curves for shifting up and down between two gears is necessary to avoid constant shifting backwards and forwards at an operating point.

6.6.4 Continuously Variable Passenger Car Transmissions The power available from an internal combustion engine cannot be fully exploited by the finite number of switching steps in traditional geared selector gearboxes. With a continuously variable transmission the engine can be operated at the ideal operating point for economy or performance as required (see also Sections 4.5 and 5.3.4).

CVT - Continuously Variable Transmission I

POWER TRANSMISSION SYSTEM I

Mechanical

Electrical

Hydraulic I

Hydrodynamic

I

I

Hydrostatic

.~H'y'SICALF>AINCIPdE: I

Friction

Guiding fluid flows

Friction gear

Torque and speed are converted to and from pressure and flow rate

>OPERATrNGiF>AINCIP~E J I

PUlley drive

I

I

... .

. . ...

Converter

Controlling current and voltage

·1

Pump/motor combination

Generator/battery supply electric motor

I

Belt I

II

Chain I I

I

Tensionallinkll Thrust link

I

Cylindrical

II

Flat

II

Tapered

II

Toroidal

fIgure 6.31. OvervIew ofCYT designs

t.L-_...,...-_ _--,--..,...--.---------------

These transmissions are referred to as CVT (Continuously Variable Transmissions). Figure 6.31 gives an overview of various CVT designs. The continuously variable transmis. (' S18n~)

.... 0'"vv

1..l

1n p'1s"pngc'r (,'1rs ~.L.,' '";.~ .

""pr! l.J-,-,-,,-,l~ .........

.L ....

',""""-d..,'

r

a1m nJ~)"t ",;tho" t

'1 p C-A (d.,""

....

"J..\...

,:'.',

··'·:.:·:.::~j:";::{"'t;;Ti':·';,::~~:j;:·:::,'··i':O:'i';"~"""";"tri_$">{""afjirflfff;';"¥=."'mw~VM;"'''*'~''~~''~''i':~'::::~';'';'~~;;::~··

.

6.7

Commercial Vehicle Transmissions

145

The total ratio i of t~is transmission is derived from iv multiplied by the ratios of the two gear stages located In the power flow. Here either the two clutches C 1 and C4 or the two clu~ches ~3 ~nd C2 are closed. The profile of the total ratio i as a function of the taper disc radIUs ratIo lV and the closed clutches is as shown in the diagram on the right of Figure 6.35.

6.7 Commercial Vehicle Transmissions For transmissions of commercial vehicles up to 4.0 t gross weight rating, the explanations in the preceding Section 6.6 "Passenger Car Transmissions" apply. The following passage relates to transmissions for commercial vehicles exceeding 4.0 t gross weight. Table 6.10 lists common types of commercial vehicle transmissions. Depending on how their idler gears are positively locked to the shafts, manual transmissions can be subdivided into

o o

non-synchronised constant-mesh transmissions and synchronised transmissions

and by the shifting system (see also Section 9.1) into

o o

direct shifting: gearshift lever at the transmission housing, indirect shifting: gearshift lever and transmission physically separated (remote shift).

Table 6.10. Market shares and appli
10 _0.8 tv" 0 Ratio spread :;:: Ii . 8 0 co if :;:: 1\ Whole transmission -=0.6 - i. co 0) a: 6 i I>< .,. > 0 Ii. I>< -oJ 0.4 f)< :i~ 4 t1

-Ii:

2

°

0.2 -

°

\

>( Lii

ii

~ i'YC: '-i~' / >~ II;:: il/ ;,!r/ tn I ••·•·•·

'/

1 234 5 6 7 8 1 2 345 6 7 8 LH LHLH LHLH LH LHLH LH LHLH LHLH LH LHLH Gear Gear Figure 6.40. Compressing and expanding the gear sequence with splitter unit and rangechange unit. Based on the example of the 16-speed commercial vehicle gearbox shown in Figure 6.44. L = Low (slow); H = High (fast)

,

,I

6.7

Commercial Vehicle Transmissions

151

This was based on a 4-speed main gearbox, with the simplifying assumption that the gear centres in each group are the same, and that for the torque multiplication in the splitter unit isplit = 1.2, in the range-change unit iR = 3.5, and in the main gearbox i G , H= 4. The characteristic value KG is determined by the number of gear pairs per range unit, an incremental torque conversion factor, and the total ratio achievable. The resultant values represent a relative measure of the physical dimensions of the transmission, and are not absolute indications of size. They serve rather to compare the transmissions with each other. The smaller the transmission characteristic value KG, the smaller the dimensions of the transmission. Front-mounted and rear-mounted range units, splitter units and range units were tried in different configurations. The results are shown in Table 6.11. The various possible configuration variants can be assessed in terms of their dimensions using this table. Table 6.11. Various combinations of range transmissions and their characteristic values KG. F Front-mounted; R Rear-mounted; S Splitter unit; R Range-change unit; Main Main gearbox

No.

Combination FS FR Main RS

x

S1 S2 R1 R2 RS1 RS2

x x

RS3 RS4

x x

x x x x x x x

x

x

RR

x x x x

x x

Gear characteristic value KG 3.17 5.63 2.89 1.93 3.15 5.74 2.27 2.05

Gearbox characteristic value KG o 1 2 345 6 I

--

I

I

I

/////////1 V////////

// //

/J

I"" " " " "'"" " I"" " " " " xx x x

x

x

The combinations reviewed can be subdivided into the following three categories:

o o o

main gearbox with splitter unit (S 1 and S2), main gearbox with range-change unit (R 1 and R2), main gearbox with splitter unit and range-change unit (RS 1 to RS4).

This reveals the combinations front-mounted splitter unit (S 1), rear-mounted rangechange unit (R2) and front-mounted splitter unit with rear-mounted range-change unit (RS4) as the most favourable in their respective categories in terms of physical dimensions. Figure 6.41 shows common configurations of two;- and three-range gearboxes. A high degree of flexibility can be achieved using the modular principle of two or three individual transmissions flanged together. Multi-range transmissions are in principle possible in passenger cars as well. In multi-range transmissions shifting times become extended, since several junctions have to be shifted in some gears. The overall shifting time should also be less than a second under unfavourable conditions. Multi-gear transmissions can be constructed with a small number of gear pairs if several junctions are shifted simultaneously when changing gear. Theoretically,

__

l~,_'

g_e_a_r_s_c_a_n.b._e_p_ro_d_u,..c_e_d_w_it,_h_p_g_e_ar--rop_ai_r_s_.

(6.6)

.

_

,

3-~ange

Main gearbox with front-mounted splitter unit and rear-mounted range-change unit

,I

x[W

gearbox

Front-mounted splitter unit isplit = 1.1-1.3

Main gearbox

Rear-mounted range-change unit iR = 3-4

Main gearbox

Rear-mounted range-change unit iR = 3-4

2-range gearbox

Main gearbox with rear-mounted range-change unit

2-range gearbox

Main gearbox with front-mounted splitter unit

Front-mounted splitter unit i split . = 1. 1-1 .3

Main gearbox

Figure 6.41. Conventional configurations with two-range and three-range gearbox

1 -,

r-

2 -,L_---:r--

3~

4~

4

i---Jr---

6~

7

=2(P-1)

z

= number of speeds

p

=number ofgear pairs

---L-

one stage

2~

3~

5

z

r-

1-,

8

r-

5

r...r---

-'L

6~ -,L_ _. r - - - -

7

r-

8

9~ 10~

11--.._ _r - -

12 13 14 15 16

~

"L--I

L-I.....---L....Jr---

Figure 6.42. Gearbox diagrams and power flows of coaxial multi-stage transmissions

6.7

153

Commercial Vehicle Transmissions

16 14 1\1

en

0-

-.... 0

.0

8

! : ,/ ,/

::l

6

2

/'

//' I :~

;~

4

Iji

/

/

/

,

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Equation 6.6 applies when all gearwheels can be shifted and each gearwheel has its own shaft. Such a transmission, in which the power is transmitted several times from one shaft to the other in individual gears, is also known as a multi-staRe transmission. In multistage transmissions the ranges are reduced to individual gear pairs. In addition to the high level of engineering complexity for the shaft junctions, several junctions have to be shifted at the same time when changing gear. Multi-stage transmissions with up to 5 gear pairs and 16 gears are shown in Figure 6.42. Shifting two or more junctions at the same time can lead to high shifting times. Depending on the design, the following transmissions can be constructed using p = 6 gear pairs (Figure 6.43):

o single-range countershaft transmission, Fig. 6.43/1: o multi-stage transmission, Figure 6.43/2:

6 forward gears 32 forward gears o two-range splitter gearbox, Figure 6.43/3.1: 10 forward gears o two-range range gearbox, Figure 6.43/3.2: 8 forward gears o three-range splitter/range gearbox, Fig. 6.43/3.3: 12 forward gears The multi-stage transmission is of no practical interest because of the many junctions to be shifted. If a splitter unit and a range-change unit are combined with a 4-speed main. gearbox, this results is a 16-speed transmission. Here the ratio spreads of the three groups are selected in such a way that all 16 selectable combinations of gear steps are arranged in steps useful for the driver (Section 4.4.2 "Geometrical Gear Steps") (Figure 6.40c).

6.7.3 Practical Design of Two- and Three-Range Transmissions The normal designs are two- and three-range transmissions with up to 16 gears (2 x 4 x 2) [6.17] (see also Figure 4.2). A larger number of gears is in principle possible, but in practice no longer relevant since it involves excessively frequent gearshifting by the driver. The ZF 16 S 109 16-speed commercial vehicle transmission (Figure 6.44) can serve as an example of a 16-speed three-range type gearbox. The main gearbox is a 4-speed countershaft transmission. The two gear pairs of the countershaft type two-speed splitter unit are located on the transmission input side. A planetary type two-speed range-change unit is connected on the transmission output side. The following principle applies to the design of commercial vehicle transmissions: The transmission must be designed in such a way that the largest possible number of gear pairs is acted on with a small change in ratio, and the smallest possible number of gear pairs with a high change in ratio. The planetary design of the range-change unit in particular ensures compactness, bearing in mind that the range-change unit must have a large gear step, which is easy to achieve in a planetary design. The short overall length also ensures minimum shaft deflection in range-change units subject to high torque. The countershaft transmissions discussed heretofore had only one countershaft located in the power flow. The transmission diagram shown in Figure 6.45 of the Eaton Twin Splitter transmission has two countershafts both for the 4-speed main gearbox and for the 3-speed rear-mounted splitter unit. The power transmitted is split between both countershafts, and flows back to the main transmission shaft. The power split enables the gearwheels to be approximately 40% narrower than in a conventional countershaft transmission. The transmission is physically shorter, but wider. Short transmissions are advantageous especially in tractors. The shorter the transmission, the more favourable the proportions (the deflection angle resulting from the vertical offset and the longitudinal distance to the final drive) for the propeller shaft connected to the transmission.

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6.7

Commercial Vehicle Transmissions

155

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In order to ensure uniform loading (load compensation) of the gearwheels in both branches of the power splitter, the main shaft does not run in radial bearings but is merely radially guided. It is centred between the two countershafts when under load. Since the main shaft is not capable ot absorbing large axial forces, straight cut spur gears arc used. To still achieve good running characteristics, gearwheels with a high contact ratio c> 2.2 and high contact gearing are used.

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The main gearbox of the twin splitter has four forward gears and one reverse gear. The rear-mounted splitter unit has three gears: one direct gear iD = 1.0, one speed-increasing gear is, H = 0.79 and one speed-reducing gear is, L = 1.28. This gives 12 forward gears. Rear-mounted splitter units are not usually used, because of the face widths required (see also Table 6.11). It is nevertheless used in this case because of the low overall face width resulting from the power split. The main gearbox is constant mesh, the rear-mounted splitter unit is synchronised. The transmission is available as a semi-automatic. Figure 6.46 shows the gearbox diagram of the Fuller (Eaton) RT 9513 Roadranger transmission. It also has two countershafts. The range-change unit is at first sight very similar to the Twin· Splitter transmission. But in the Roadranger transmission the rangechange unit is in the form of a combined splitter unit/range-change unit (is, H = 0.87; is, L = iD = 1.0; iR = 3.38). CG

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Figure 6.46. Gearbox diagram: Roadranger (Fuller) 4 x 3 + crawler = 13 speeds. 2-countershaft gearbox, 3-speed rear-mounted splitter/range-change unit; CG Constant gear SH Rear-mounted splitter unit high D Direct R Range

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6.7

Commercial Vehicle Transmissions

157

The main gearbox of the Roadranger transmission has five forward gears and one reverse gear, with the first gear being used only once in combination with the range-change unit as a crawler gear. The other four forward gears are combined with the splitter unit. This results in 13 forward gears for this transmission. The reverse gear can also operate with range-change unit, providing two reverse gears.

6.7.4 Semi-Automatic Manual Commercial Vehicle Transmissions In line with the explanations in Section 6.2.2 the designation "semi-automatic" relates to the operations "Engaging the clutch/Moving off' and "Changing gear". In semi-automatic transmissions one of these processes is automatic. This results in the breakdown shown in Table 6.12 listing the various degrees of automation of transmissions from manual gearboxes (automation level 0) through to fully automatic transmissions (automation level 4). Table 6.12. Degrees of automation of manual passenger car and commercial vehicle gearboxes Degree of automation

Method of engaging drive

Shifting clutch action

Gear selection method

0

Foot-activated master clutch

Foot-activated clutch operation

Manual activation of a shift lever

1

Foot-activated master clutch

Automated clutch operation

Manual activation of a shift lever

2

Automated master clutch

Automated clutch operation

Manual activation of a shift lever

3

Automated master clutch

Gear change initiated by foot-activated clutch operation

Manual gear pre-selection from keypad

4

Automated master clutch

Automated clutch operation

Automated gear selection and engine management

In automation level 2 the driver just engages the desired gear by activating the gearshift lever, then the engaging action and moving off take place automatically.. One e:ample of this design is the ZF torque converter clutch shown in Figure 6.47, combmed WIth the 16speed transmission in Figure 6.44.

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·tf figure G.47. Gearbox diagram of (l 16-speed commercial vehicle gearbox with torque g

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6.7.5 Fully Automatic Commercial Vehicle Transmissions In fully automatic commercial vehicle gearboxes, both engaging the clutch/moving off and changing gear are automated. The following designs can be distinguished:

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o o

conventional 3-speed to 7-speed automatic transmissions (consisting of torque converter and rear-mounted planetary gear), fully automatic commercial vehicle countershaft transmissions, continuously variable transmissions.

See also Section 6.6.3 "Fully Automatic Passenger Car Transmissions". Conventional automatic transmissions have powershift. The electronically controlled commercial vehicle countershaft transmissions have shifting with power interruption. Countershaft transmissions have the advantage of being based on conventional range gearboxes and being able to provide up to 16 speeds with acceptable production engineering requirements. By comparison, conventional automatic transmissions are available with a maximum of 7 speeds.

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